Working vehicle and vehicle speed control method thereof, variable power engine and power setting method thereof, and vehicle with variable power engine and power control method thereof

ABSTRACT

In a working vehicle, an electronic control regulating valve  121  controls a connection force in clutch mechanisms  35  and  36  in a transmission  30 , and a input side revolution signal of the transmission, an output side revolution signal, a speed stage position signal and a pedal angle signal of an inching pedal  114  are inputted to control a slip in the transmission  30 . Further, a pedal angle signal of an accelerator pedal  107  is inputted to cope with a hyper-slow mode. Furthermore, in an engine, a charged pressure is controlled in stages to carry out an appropriate power control suitable for a number of stages of the transmission  30.

BACKGROUND OF THE INVENTION

1. Field of the Invention

The present invention relates to a construction machine such as a motorgrader, a rough terrain crane, a power shovel, a wheel loader, abulldozer and others, or a working vehicle such as a dump truck, aforklift and others, and a speed control method thereof. Further, thepresent invention relates to a variable power engine capable of changingits power and a power setting method thereof, the engine being used inthe above vehicles. Furthermore, the present invention relates to avehicle with a variable power engine such as described above and a powercontrol method thereof.

2. Description of the Related Art

Recently, in a working vehicle, e.g., a machine such as a motor grader,a gear-contained transmission with an electronically controlled clutchhaving a gear train and a plurality of hydraulic (fluid pressure)clutches, combination of which can be changed in various ways, is used,and the gear-contained transmission with the clutch controlled by asolenoid operated valve to carry out various operations is increasing.

(Inching Pedal)

In a motor grader, a special work such as leveling the ground by using ablade while turning a corner may be performed. In such a case, a numberof revolutions of the engine must be increased in some measure and, inthis high velocity revolution, the vehicle speed is also increased fromthe nature of the case, which makes it difficult to turn a comer whilecarrying out the necessary work. Therefore, there is a request forreducing the vehicle speed while maintaining the high velocityrevolution.

In order to satisfy such a request, a foot-operated inching pedal(clutch pedal) is provided, and stepping on this inching pedal can slidea predetermined clutch of the transmission to decrease the speed,thereby filling the needs of a necessary work and turning a corner andthe like.

In case of the working vehicle, when starting the machine fromstandstill, since starting is possible from a relatively high speedstage as different from a general vehicle, the machine is started at aplurality of speed stage positions. When starting at different speedstage positions in this manner, since an input shaft is different froman output shaft in speed reducing (speed increasing) ratio for eachspeed stage position, the connection state of the clutch differs eventhough the inching pedal is stepped on at a similar angle. In otherwords, when the inching pedal stepping angle is unchanged, the machinecan be started and travel at a desired speed at a low speed stageposition, whilst the machine can not be started or travel at a lowerspeed than expected at a higher speed stage position. Therefore, thereis a problem that a desired speed can not be obtained unless the inchingpedal stepping angle is reduced (unless the pedal is released) at thehigher speed stage position.

As a technique for equaling the feel of stepping on the inching pedalwhich differs in accordance with each speed stage position, there is aninvention disclosed in Japanese patent laid-open publication No. Hei10-246318.

In the invention disclosed in this publication, a solenoid workingpressure regulating valve is provided to each clutch constituting eachspeed stage of the transmission; one solenoid operation manualregulating valve (proportional valve) is provided to some regulatingvalves out of the above pressure regulating valves on the upstream side;and each valve is controlled by a controller in accordance with aninching pedal stepping angle and a speed stage position, therebycontrolling the connecting force of the clutch to be controlled inaccordance with a change in the speed stage (selected gear ratio).

According to the invention disclosed in this publication, since theconnecting force of the clutch is changed in accordance with the speedstage position, the inching pedal stepping position can be roughlyassociated with the speed of the machine irrespective of the speed stageposition of the transmission, but an input number of revolutions and anoutput number of revolutions of the transmission are not monitored atall, and the sufficient (detailed) control of the output shaft, i.e.,the vehicle speed can not be necessarily performed.

Further, since the pressure regulating valve of each clutch iscontrolled by one proportional valve provided on the upstream sidethereof, there is also such a problem as that the detailed control ofeach clutch is not necessarily adequate.

(Hyper-slow Mode)

In the motor grader, a work for finishing the road surface or the bankwith the extremely high accuracy may be carried out. In such a case,there is a request for realizing the very low speed of, e.g.,approximately 1.0 km/h from the necessity of the work.

On the other hand, as to a number of revolutions of the engine, the highspeed revolutions in some measures is required from the necessity of thework and, in such a high speed revolution, the vehicle speed isnecessarily increased, which makes it difficult to realize the very lowspeed running while carrying out the necessary work. Therefore, there isa need for decreasing the vehicle speed while maintaining the high speedrevolution.

In order to satisfy the above needs, there have been proposed manyvariable displacement torque converters, i.e., torque converters capableof automatically converting attitudes or shapes of three vanes, namely,an impeller, a turbine vane and a stator of the torque converter inaccordance with a range of the high and low speeds.

With the working vehicle having the above-described variabledisplacement torque converter, running at a very low speed and apredetermined finishing operation can be efficiently carried out.

However, since the variable displacement torque converter has thecomplicated structure, it is very expensive, resulting in an increase inthe price of the vehicle itself.

Further, the rising interest in the recent environmental concerns leadsto an issue of noise reduction of the working vehicle.

That is, a number of steady-state revolutions of a conventional engineis determined to be approximately 2,500 RPM and a number of revolutionsin the low idling is determined to be approximately 800 RPM.

On the other hand, in view of the noise reduction, a number ofsteady-state revolutions of approximately 2,000 RPM and a number ofrevolutions in the low idling of 800 RPM are required. This brings upthe following new problem when realizing the very low running such asdescribed above.

That is, if a forward first speed in the first in forward in aconventional number of steady-state revolutions is set to, for example,3.45 km/h, the vehicle speed in the low idling becomes 1.1(=3.45×800/2,500) km/h, which is the hyper-slow running with which thefinishing work is possible.

Meanwhile, if a number of steady-state revolutions is set to 2,000 RPMwithout changing the structure of the transmission, the vehicle speed inthe low idling becomes 1.38 (=3.45×800/2,000) km/h.

However, the vehicle speed of 1.38 km/h is too high for the finishingwork using the motor grader.

Thus, a working vehicle having a simplified structure by which a numberof steady-state revolutions of the engine can be reduced and the verylow speed running is also enabled is desired.

(Variable Power Engine)

The following can be true to the above-described engine of a workingvehicle.

An engine power is usually set so as to obtain a given fixed maximumpower. However, there is a request for changing a set value of a maximumpower in accordance with circumstances depending on vehicles andmachines.

For example, in case of the construction machine, a large power directlyrelates to the high work efficiency. However, in a low speed range,since the speed decreasing ratio of the transmission is large, a largepower may cause the drive force to exceed a road surface adhesivecoefficient of a wheel or a crawler and, in such a case, the wheel orthe crawler go into a slide. Slippage of the wheel and the like leads todifficulty in the work and the deteriorated drive controllability, andwear-out of the wheel and the like is caused to shorten the life,thereby degrading the work efficiency.

In order to prevent this, an appropriate power is determined inaccordance with the speed stages of the transmission in a regular workarea, for example. Thus, the engine can afford to output, but the enginepower is suppressed in order to prevent slippage during the work in thelow speed stage even though the highly efficient work is possible withmore power when working in the higher speed stage.

In the above-described case, if the power is variable, an appropriatepower can be selected in the work, thereby improving the workefficiency.

Therefore, there has been developed an engine (variable power engine)with which an engine power is variable by controlling fuel oilconsumption. FIG. 31 indicates an example of running and rim pullperformances of a general vehicle with a variable power engine. FIG. 31indicates the relationship between a vehicle speed (horizontal axis) anda rim pull (vertical axis) of a vehicle having a six-speed change gear.The characteristic indicated by a solid line represents the character ofall speeds from the first speed to the sixth speed in the state of theusual power (normal power), and the characteristic indicated by a brokenline represents the characteristic of the high speed range from thethird speed to the six speed in the high power state.

In such a vehicle, the land clearing work is conducted at the secondspeed; the work for scattering a material such as ballast, the work forleveling the gravel path and the light load work such as snow removalare carried out at the third to fifth speeds; and the running or lighterload work is performed at the sixth speed.

In this case, since slippage of the wheel and the like does not occurwith the high power in the light load work or running at the third orhigher speed, the engine power is set to the high power characteristicindicated by the broken line to conduct the highly efficient work.

There are currently known several methods for obtaining a variablepower. For example, there are the following methods.

1) An electronically controlling method adopting electron governorcontrol.

2) A method disclosed in U.S. Pat. No. 4,785,778. An apparatus used inthis method has such a structure as that one end of a main governorspring is brought into contact with a movable rack controlling a fueloil consumption of a fuel injection system while the other end of themain governor spring is associated with a control lever and a positionof the control lever is changed by a hydraulic cylinder. When theposition of the control lever is displaced by using the hydrauliccylinder in a direction for thrusting the main governor spring or in theopposed direction, a spring pressure is changed to control the fuel oilconsumption.

3) A controlling method using a boost compensator. In this method, asupercharged pressure of a turbo charger to be applied to a boostcompensator is subjected to on/off control by using an electromagneticvalve in an engine provided with a fuel injector having the boostcompensator, and a high power state in which the supercharged pressureis supplied and a normal power state in which no supercharged pressureis supplied are set to obtain the two-stage engine power.

Regarding to each method described above, in the electronic governor of1), a flexibility of the power setting is high but the price is high.Further, attachment on the spot (outdoor and field) and general repairby a serviceman in case of a failure are difficult.

As to the method 2) disclosed in the US patent, the system has arelatively simple structure, and repair is easy, but on-the-spotretrofitting is difficult after distributed to a customer.

In the control 3) using the boost compensator, the system is inexpensiveand simple, repair is easy and retrofitting is possible. However, theflexibility of the power setting is low, and since the superchargedpressure is subjected to only the on/off control, the intermediate powersetting is impossible even if such setting is necessary. In addition,when turning off supply of the supercharged pressure to the boostcompensator, since a duct connecting the turbocharger to the boostcompensator is simply opened in the atmosphere, the air supercharged bythe turbocharger is partially wastefully discharged in the atmosphere.

In a vehicle having a recent variable power engine, a torque converterprovided between the engine and the transmission is often seen. However,in the vehicle having the torque converter, although the drivingoperation is facilitated, slippage in the torque converter occurs todecrease the mechanical efficiency of power transmission, and reductionis the efficiency during driving at the high speed stage becomes aproblem in particular.

Accordingly, a torque converter having a lockup mechanism for directlycoupling (locking up) the torque converter at the high speed stage wasdeveloped, and a motor grader provided with the torque converter havingsuch a lockup mechanism is on sale from CATERPILLAR in US (see thecatalog “24H, Motor Grader”, p. 192, 1996, CATERPILLAR).

However, the motor grader of CATERPILLAR does not take the relationshipbetween the variable power engine and the lockup mechanism into accountat all, and the satisfactory driving control is not realized.

Further, in order to make the running performance of the vehicle havingthe variable power engine further complete, an all wheel drive (AWD)type vehicle in which a driving wheel can be switched to all the wheelsis on sale. This type of all wheel drive type motor grader is on marketfrom JOHN DEERE in US (see the catalog “Motor graders 770C, 770CH,772CH”, 1997, JOHN DEERE).

Although the motor grader of JOHN DEERE takes the relationship betweenthe variable power engine and the all wheel drive into consideration, itdoes not include the torque converter or that having the lockupmechanism for improving facilitation of the running operation. Thismotor grader does not sufficiently enhance the operability and thecontrollability of the vehicle having the variable power engine.

SUMMARY OF THE INVENTION

An object of the present invention is to provide a working vehicle withwhich an operator can work without changing his/her feeling irrespectiveof a speed stage position in a transmission.

Another object of the present invention is to provide a working vehiclecapable of performing detailed control in each speed stage position in atransmission.

Still another object of the present invention is to provide a workingvehicle capable of selecting a degree of connection of a clutchmechanism by operating an inching pedal according to an operator'spreference.

In order to achieve the above aim, according to the present invention, adegree of connection (slip state) of a clutch mechanism by operating aninching pedal is controlled by monitoring an angle of an inching pedal,a speed stage position (shift position) of a transmission, and numbersof revolutions of the transmission on input and output sides.

Further, another objection of the present invention described above isattained by connecting an electronic control regulating valve (ECMV) toeach clutch of the transmission.

Furthermore, still another object of the present invention describedabove is attained by adding a mechanism capable of selecting a degree ofconnection of the clutch mechanism by operating the inching pedalaccording to an operator's preference.

Specifically, the present invention defined in the first embodimentprovides a working vehicle comprising: an engine; a gear-containedtransmission with a clutch having a plurality of clutch mechanisms forconverting revolutions of said engine into speed ratios of multiplestages and a gear train; a connection force control mechanism forcontrolling connection force in said clutch mechanism in saidtransmission; an inching pedal operated in order to generate a slip in apredetermined clutch mechanism of said transmission; an input siderevolution detection mechanism for detecting a number of revolutions onan input side of said transmission to output an input side revolutionsignal; an output side revolution detection mechanism for detecting anumber of revolutions on an output side of said transmission to outputan output side revolution signal; a speed stage position detectionmechanism for detecting a speed stage position of said transmission tooutput a speed stage position signal; a pedal angle detection mechanismfor detecting a stepping angle of said inching pedal to output a pedalangle signal; and a transmission controller to which said input siderevolution signal of said input side revolution detection mechanism,said output side revolution signal of said output side revolutiondetection mechanism, said speed stage position signal of said speedstage position detection mechanism and said pedal angle signal of saidpedal angle detection mechanism are inputted and which outputs aconnection force control signal to said connection force controlmechanism based on said input side revolution signal, said output siderevolution signal, said pedal angle signal and said speed stage positionsignal in such a manner that a number of revolutions on said output sidein said transmission becomes a predetermined value.

According to the present invention, since a degree of connection (slipstate) of the clutch mechanism is controlled by operating the inchingpedal while monitoring an angle of the inching pedal and a speed stageposition (shift position) of the transmission as well as numbers ofrevolutions of the transmission on the input and output sides, theclutch mechanism can be appropriately connected irrespective of thespeed stage position.

In the present invention, the term “speed stage position” is used as aconcept including forward and reverse directions. Further, the inputside and the output side of the transmission do not necessarily indicatean input and an output of the transmission itself but means the sideswhich the clutch mechanism generating slippage is engaged to or releasedfrom. That is, if the transmission has a plurality of shafts and theclutch mechanism is provided to each shaft, when slippage is generatedin one or more clutch mechanisms associated with a first shaft and asecond shaft for example, the first shaft is on the input side and thesecond and the subsequent shafts are on the output side. Therefore, anumber of revolutions on the output side can be detected by using thesecond and the subsequent shafts. This idea can be applied to detectionof a number of revolutions on the input side when the shaft associatedwith slippage is the second or the subsequent shafts. That is, theupstream side of the clutch mechanism associated with slippagecorresponds to the input side and the downstream side of the same to theoutput side.

In the working vehicle according to the first embodiment, the presentinvention provides a working vehicle, wherein said controller isprovided with: an actual speed ratio calculating function forcalculating an actual speed ratio in said transmission from a ratio ofsaid input side revolution signal and said output side revolutionsignal; a target speed ratio calculating function for calculating atarget speed ratio in said transmission from said pedal angle signal;and a control signal transmitting function for outputting saidconnection force control signal to said connection force controlmechanism from a difference between said target speed ratio and saidactual speed ratio and said speed stage position signal in such a mannerthat said actual speed ratio becomes a predetermined value.

According to the present invention, the simple calculation in thecontroller can control slippage of the clutch mechanism.

In general, although the speed ratio means a ratio of a number ofrevolutions on the input side to a number of revolutions on the outputside, in the present invention, basically, the speed ratio may bedetected as a slip ratio since detection of the anteroposterior slipstate of the clutch mechanism generating slippage is sufficient.

In the working vehicle according to the first embodiment, the presentinvention provides a working vehicle, wherein said controller isprovided with: an actual speed ratio calculating function forcalculating an actual speed ratio in said transmission from a ratio ofsaid input side revolution signal and said output side revolutionsignal; a target speed ratio calculating function for calculating atarget speed ratio in said transmission from said pedal angle signal andsaid speed stage position signal; and a control signal transmittingfunction for outputting said connection force control signal to saidconnection force control mechanism from said target speed ratio and saidactual speed ratio in such a manner that said actual speed ratio becomesa predetermined value.

According to the present invention, the advantages similar to those ofthe first embodiment can be obtained.

A difference between the target speed ratio and the actual speed ratiois previously calculated to perform correction based on the speed stageposition signal and an obtained correction signal is outputted from thecontroller to the connection force control mechanism as the connectionforce control signal in the invention described in the first embodiment,whilst correction is carried out based on the speed stage positionsignal in advance when calculating the target speed ratio and the targetspeed ratio incorporating the content of the speed ratio position signalis compared with the actual revolution signal to output the connectionforce control signal to the connection force control mechanism in thepresent invention.

In the working vehicle according to the first embodiment, the presentinvention provides a working vehicle, wherein said connection forcecontrol mechanism is an electronic control regulating valve (ECMV) whichis coupled to a predetermined clutch mechanism to be controlled among aplurality of said clutch mechanisms in said transmission and controls anamount of working fluid to said clutch mechanism in response to saidconnection force control signal from said controller.

According to the present invention, a working fluid quantity to theclutch mechanism can be controlled by the electronic control regulatingvalve (ECMV) capable of accurate control.

In the working vehicle according to the first embodiment, the presentinvention provides a working vehicle, wherein said electronic controlregulating valve includes: a pressure control valve to which saidconnection force control signal from said controller is applied andwhich converts a pressure into a fluid pressure responsive to saidsignal; and a flow rate detection valve operated by a hydraulic pressuresignal from said pressure control valve.

According to this invention, the further accurate control is possible.

In the working vehicle according to the first embodiment, the presentinvention provides a working vehicle, wherein said gear-containedtransmission with said clutch includes a plurality of directionswitching clutch mechanisms and a plurality of speed switching clutchmechanisms and a clutch mechanism for generating a slip in accordancewith an operation amount of said inching pedal is said directionswitching clutch mechanism.

In general, although the clutch mechanism for generating a slip requiressatisfactory cooling means because of a large calorific value, thecooling means having the large cooling power may be provided only to thedirection switching clutch mechanisms whose number is relatively smallaccording to the present invention, thereby reducing the manufacturingcost.

In the working vehicle according to the first embodiment, the presentinvention provides a working vehicle, wherein said controller isdesigned to have a control function such that a number of revolutions onsaid output side becomes within a predetermined deviation.

According to the present invention, since control is carried out with agiven allowance, hunching does not occur during the control.

In the working vehicle according to the first embodiment, the presentinvention provides a working vehicle, wherein to said controller isconnected to a characteristic change mechanism by which a content ofsaid connection force control signal to be outputted to said connectionforce control mechanism can be changed in accordance with a workingcondition, an operator's preference and others.

According to the present invention, the slip state of the clutchmechanism can be appropriately changed in accordance with differences inthe work condition or an operator's preference.

Yet another object of the present invention is to provide a workingvehicle and a speed control method thereof capable of inexpensivelyobtaining appropriate working power and hyper-slow running with a simplestructure.

The present invention intends to attain the above aim by making ajudgement upon whether the hyper-slow running is carried out by usingthe controller and controlling the connection force of the clutchmechanism in the transmission.

Specifically, the present invention described in the first embodimentprovides a working vehicle comprising: an engine; a gear-containedtransmission with a clutch having a plurality of clutch mechanisms forconverting revolutions of said engine into speed ratios of a pluralityof stages and a gear train; a connection force control mechanism forcontrolling connection force in said clutch mechanism in saidtransmission; an output side revolution detection mechanism fordetecting a number of revolutions on an output side of said transmissionto output an output side revolution signal; a speed stage positiondetection mechanism for detecting a speed stage position of saidtransmission to output a speed stage position signal; an acceleratorpedal operated in order to increase a number of revolutions of saidengine; an accelerator pedal angle detection mechanism for detecting astepping angle of said accelerator pedal to output an accelerator pedalangle signal; and a transmission controller to which said output siderevolution signal of said output side revolution detection mechanism,said speed stage position signal of said speed stage position detectionmechanism and said accelerator pedal angle signal of said acceleratorpedal angle detection mechanism are inputted, which makes a judgmentupon whether a current mode is a hyper-slow running mode based on saidoutput side revolution signal, said speed stage position signal and saidaccelerator pedal angle signal, and which outputs a connection forcecontrol signal to said connection force control mechanism in such amanner that a number of revolutions on said output side in saidtransmission becomes a predetermined hyper-slow value in case of saidhyper-slow running mode.

According to the present invention, since the connection force of theclutch mechanism of the transmission is controlled by using the controlsignal from the controller through the connection force controlmechanism, the hyper-slow running can be acquired with a simplestructure.

In the present invention, the term “speed stage position” is used as aconcept including the both forward and reverse directions.

In the working vehicle according to the first embodiment, the presentinvention provides a working vehicle, wherein said controller isprovided with: a hyper-slow running mode judging function whichdetermines said hyper-slow running mode when a vehicle speed calculatedfrom said output side revolution signal is smaller than a predeterminedspeed, e.g., 2.0 km/h or 1.8 km/h larger than a hyper-slow target speed,e.g., 10 km/h, said accelerator pedal angle signal is in a standby modeand said speed stage position signal indicates a predetermined low speedstage position; and a control signal transmitting function foroutputting said connection force control signal to said connection forcecontrol mechanism in such a manner that a vehicle speed set in saidhyper-slow running mode is obtained when said hyper-slow running mode isdetermined by said hyper-slow running mode judging function.

According to the present invention, the slippage of the clutch mechanismcan be controlled by simple calculation in the controller, therebyobtaining the necessary hyper-slow running.

In the working vehicle according to the first embodiment, the presentinvention provides a working vehicle, wherein said connection forcecontrol mechanism is an electronic control regulating valve (ECMV) whichis coupled to a predetermined clutch mechanism to be controlled among aplurality of said clutch mechanisms of said transmission and whichcontrols an amount of working fluid to said clutch mechanism in responseto said connection force control signal from said controller.

According to the present invention, the working fluid quantity to theclutch mechanism can be controlled by the electronic control regulatingvalve (ECMV) capable of accurate control.

In the working vehicle according to the first embodiment, the presentinvention provides a working vehicle, wherein said electronic controlregulating valve includes: a pressure control valve to which saidconnection force control signal from said controller is applied andwhich converts a pressure into a fluid pressure responsive to saidsignal; and a flow rate detection valve operated by a hydraulic pressuresignal from said pressure control valve.

According to the present invention, the further accurate control ispossible.

In the working vehicle according to the first embodiment, the presentinvention provides a working vehicle, wherein said gear-containedtransmission with said clutch includes a plurality of directionswitching clutch mechanisms and a plurality of speed switching clutchmechanisms and a clutch mechanism whose clutch connection force iscontrolled for generating a hyper-slow speed in response to saidconnection control signal from said connection force control mechanismis said direction switching clutch mechanism.

In general, although the clutch mechanism generating a slip requiressatisfactory cooling means because of a large calorific power, provisionof the cooling means having a large cooling power to only the directionswitching clutch mechanisms whose number is relatively small can sufficethe invention, thereby reducing the manufacturing cost.

In the working vehicle according to the first embodiment, the presentinvention provides a working vehicle, wherein said controller isdesigned to have a control function such that a number of revolutions onsaid output side becomes within a predetermined deviation.

According to the present invention, since control is performed with apredetermined allowance, hunching does not occur during the control.

The present invention in the second embodiment provides a workingvehicle comprising: an engine; a gear-contained transmission with aclutch having a plurality of clutch mechanisms for convertingrevolutions of said engine into speed ratios of a plurality of stagesand a gear train; a connection force control mechanism for controllingconnection force in said clutch mechanism of said transmission; a speedmode setting mechanism which can switch a vehicle speed to a normalrunning mode and a hyper-slow running mode and outputs a running modesignal; and a transmission to which said running mode signal is inputtedfrom said speed mode setting mechanism and which outputs a connectionforce control signal to said connection force control mechanism in sucha manner that a number of revolutions on an output side in saidtransmission becomes a predetermined hyper-slow value when said runningmode signal indicates said hyper-slow running mode.

According to the present invention, the speed mode can be appropriatelychanged by the speed mode setting mechanism.

The present invention defined in the second embodiment provides avehicle speed control method for a working vehicle comprising: anengine; a gear-contained transmission with a clutch having a pluralityof clutch mechanisms for converting revolutions of said engine intospeed ratios of a plurality of stages and a gear train; a connectionforce control mechanism for controlling connection force in said clutchmechanism of said transmission; an accelerator pedal operated in orderto increase revolutions of said engine; an output side revolutiondetection mechanism for detecting a number of revolutions of saidtransmission to output an output side revolution signal; a speed stageposition detection mechanism for detecting a speed stage position ofsaid transmission to output a speed stage position signal; anaccelerator pedal angle detection mechanism for detecting a steppingangle of said accelerator pedal to output an accelerator pedal anglesignal; and a transmission controller to which said output siderevolution signal of said output side revolution detection mechanism,said speed stage position signal of said speed stage position detectionmechanism and said accelerator pedal angle signal of said acceleratorpedal angle detection mechanism are inputted, which make a judgment uponwhether a current mode is a hyper-slow running mode from said outputside revolution signal, said speed stage position signal and saidaccelerator pedal angle signal and which outputs a connection forcecontrol signal to said connection force control mechanism in such amanner that a number of revolutions on said output side of saidtransmission becomes a predetermined hyper-slow value in case of saidhyper-slow running mode, wherein said clutch mechanism and said geartrain of said transmission are constituted as a plurality of directionswitching clutch mechanisms and a plurality of speed switching clutchmechanisms, and said direction switching clutch mechanism is constitutedas a forward low speed clutch mechanism and a reverse clutch mechanism,and wherein the control by said controller in said hyper-slow runningmode is effected by supplying a predetermined working fluid pressure toeither or both of any clutch mechanism on a driving side of said forwardlow speed clutch mechanism and said reverse clutch mechanism and aclutch mechanism on an opposite side among said direction switchingclutch mechanisms.

According to the present invention, since a predetermined working fluidpressure is supplied to both the driving side clutch mechanism and theopposite side clutch mechanism of either the forward low speed clutchmechanism or the reverse clutch mechanism in the direction switchingclutch mechanism to be controlled, the both clutch mechanism arebalanced to carry out the speed control, thereby easily realizing thehyper-slow running.

In the vehicle speed control method of a working vehicle according tothe second embodiment, the present invention provides a vehicle speedcontrol method for a working vehicle, wherein said control by saidcontroller in said hyper-slow running mode comprises the steps of: (1)supplying a low working fluid pressure of a first stage to said bothclutch mechanisms when a vehicle speed is a control target value orwithin a deviation obtained by adding a predetermined difference to saidcontrol target value; (2) maintaining, on one hand, supply of said lowworking fluid pressure of said first stage to said clutch mechanism onsaid driving side, and, on the other hand, increasing said working fluidpressure to said clutch mechanism on an opposite side of said drivingside to serve as a braking force when said vehicle speed is higher thansaid control target value or said deviation obtained by adding saidpredetermined difference to said control target value; and (3)maintaining, on one hand, supply of said low working fluid pressure ofsaid first stage to said clutch mechanism on an opposite side of saiddriving side and, on the other hand, increasing said working fluidpressure to said clutch mechanism on said driving side to serve as aspeed increasing force when a vehicle speed is lower than said controltarget value or said deviation obtained by adding said predetermineddifference to said control target value.

According to this invention, the hyper-slow running can be produced moreaccurately

In the vehicle speed control method of a working vehicle according tothe second embodiment, the present invention provides a vehicle speedcontrol method of a working vehicle, wherein said control by saidcontroller in said hyper-slow running mode comprises the steps of: (1)supplying a low working fluid pressure for generating a slip in saidclutch mechanism on said driving side when a vehicle speed becomes notmore than a predetermined value larger than a control target value; and(2) causing a braking force to act as an appropriate value by increasingor decreasing said working fluid pressure having a predetermineddifference in accordance with a fixed cycle to said clutch mechanism onan opposite side of said driving side when a vehicle speed is higher orlower than said control target value beyond a predetermined difference,and shifting from said hyper-slow running mode to said normal runningmode when a vehicle speed becomes not less than a predetermined valuelarger than said predetermined value greater than said control targetvalue for entering said hyper-slow running mode.

According to the present invention, the hyper-slow running can befurther accurately realized.

In the vehicle speed control method of a working vehicle according tothe second embodiment, the present invention provides a vehicle speedcontrol method for a working vehicle, wherein said control by saidcontroller in said hyper-slow running mode comprises the steps of: (1)supplying a low working fluid pressure for generating a slip in saidclutch mechanism on said driving side when a vehicle speed becomes notmore than a predetermined value larger than a control target value; and(2) causing a braking force to act as an appropriate value by increasingor decreasing a working fluid pressure suitable for a deviation betweensaid actual vehicle speed and said target vehicle speed to said clutchmechanism on an opposite side of said driving side when a vehicle speedis higher or lower than said control target value beyond a predetermineddifference, and shifting from said hyper-slow running mode to saidnormal running mode when a vehicle speed becomes not less than apredetermined value larger than said predetermined value greater thansaid control target value for entering said hyper-slow running mode.

According to the present invention, the hyper-slow running can befurther accurately realized.

A further object of the present invention is to provide an inexpensivevariable power engine which has a simple structure and facilitatesrepair and retrofitting and which has a relatively high flexibility ofpower setting, and a power setting method thereof.

In order to attain the above object, the present invention provides avariable power engine, wherein a pressure state switching mechanismwhich can switch between and supply a reference pressure to be suppliedto a boost compensator and at least a first-stage predetermined pressuredecreased to be lower than the reference voltage is provided in amanifold for supplying a charged pressure (output side pressure) of aturbocharger to a fuel injector having a boost compensator, and thepressure state switching mechanism is switched in accordance with aspeed stage position signal of a transmission.

In order to attain the above object, the present invention provides apower setting method of a variable power engine, wherein the chargedpressure of the turbocharger is supplied to the boost compensator as atleast a two-or-more-stage predetermined pressure which is equal to orlower than the charged pressure and higher than an atmospheric pressurein accordance with the speed stage of the transmission.

Specifically, the present invention according to the modification of thesixth embodiment provides a variable power engine comprising: aturbocharger for supplying a charged pressure to an engine; a fuelinjector with a boost compensator for adjusting a fuel oil consumptionin accordance with said charged pressure of said turbocharger; atransmission for changing output revolutions of said engine to aplurality of speed stages; a manifold for supplying said chargedpressure of said turbocharger to said boost compensator; a pressurestate switching mechanism which is provided to said manifold and capableof switching a pressure to be supplied to said boost compensator topredetermined pressures of at least two stages equal to or lower thansaid charged pressure of said turbocharger and higher than anatmospheric pressure; and a speed stage position detection mechanism fordetecting a speed stage position of said transmission to supply a speedstage position signal to said pressure state switching mechanism, saidpressure state switching mechanism being caused to perform pressurestate switching operation by said speed stage position signal suppliedfrom said speed stage position detection mechanism.

According to the present invention, since the supply pressure to theboost compensator can be changed to a predetermined set pressure to besupplied by the pressure state switching mechanism, an engine powermatched with the operation content can be obtained, thereby improvingthe work efficiency.

In the variable power engine according to the modification of the sixthembodiment, the present invention provides a variable power engine,wherein said pressure state switching mechanism comprises: a hydrauliccircuit device which is provided to said manifold and demonstrates aresistance action or a pressure decreasing action with respect to saidmanifold; an auxiliary manifold branched off between said hydrauliccircuit device and said boost compensator in said manifold; andswitching means which is provided to said auxiliary manifold and capableof switching between a state in which said auxiliary manifold is blockedand a state in which a pressure of said auxiliary manifold is releasedand a pressure to be applied from said turbocharger to said manifold isdecreased to be a predetermined pressure of at least one stage lowerthan said pressure and higher than an atmospheric pressure.

According to the present invention, since the auxiliary manifold isprovided to the manifold for supplying the charged pressure of theturbocharger to the boost compensator and the switching means forswitching between the state in which the auxiliary manifold is blockedand the state in which the pressure is decreased to a predeterminedpressure lower than the charged pressure, the predetermined pressure canbe set with a simple structure.

In the variable power engine according to the modification to the sixthembodiment, the present invention provides a variable power engine,wherein said hydraulic circuit device provided to said manifold isconstituted by a throttle.

According to the present invention, the hydraulic circuit device canhave a very simple structure and provided inexpensively.

In the variable power engine according to the modification of the sixthembodiment, the present invention provides a variable power engine,wherein said switching means provided to said auxiliary manifold isconstituted by a two-position selector valve capable of switchingbetween a duct blocking state and a communicating state and a throttleprovided to a duct on a slip stream side of said two-position selectorvalve in said communicating state of said two-position selector valve.

According to the present invention, the switching means can beconstituted by a simple mechanism and inexpensively provided.

In the variable power engine according to the modification to the sixthembodiment, the present invention provides a variable power engine,wherein said slip stream side of said auxiliary manifold is connected toa duct on an upstream side of said turbocharger.

According to the present invention, when supplying a pressure decreasedto be not more than the reference pressure to the boost compensator, thepressure flowing from the auxiliary branch circuit can be assuredlyflowed back to the turbocharger side, eliminating the wasteful outflowof the pressure.

In the variable power engine according to the modification to the sixthembodiment, the present invention provides a variable power engine,wherein said pressure state switching mechanism is caused to switch tosaid state for blocking said auxiliary manifold when said speed stageposition signal supplied from said speed stage position detectionmechanism is a high speed stage position signal.

According to the present invention, a light work operation at a highspeed can be assuredly automatically performed with a high power inaccordance with the speed stage position.

The present invention as disclosed in the modification to the seventhembodiment provides a variable power engine comprising: a turbochargerfor supplying a charged pressure to an engine; a fuel injector with aboost compensator for adjusting a fuel oil consumption in accordancewith said charged pressure of said turbocharger; a transmission forchanging output revolutions of said engine to a plurality of speedstages; a manifold for supplying said charged pressure of saidturbocharger to said boost compensator; a pressure state switchingmechanism which is provided to said manifold and capable of switching apressure to be supplied to said boost compensator to a pressure equal tosaid charged pressure of said turbocharger and a predetermined pressureof at least one stage lower than said charged pressure and higher thanan atmospheric pressure; and a speed stage position detection mechanismfor detecting a speed stage position of said transmission to supply aspeed stage position signal to said pressure state switching mechanism,said pressure state switching mechanism being caused to perform pressurestate switching operation by said speed stage position signal suppliedfrom said speed stage position detection mechanism.

According to the present invention, it is possible to obtain an enginepower matched with a content of work as similar to the invention definedin claim 20, thereby improving the work efficiency.

In the variable power engine according to the modification to theseventh embodiment, the present invention provides a variable powerengine, wherein said pressure state switching mechanism provided to saidmanifold is constituted by a two-position selector valve capable ofswitching said communicating state of said manifold to two positions anda pressure reducing valve provided to at least one duct on a slip streamside of said two-position selector valve.

According to the present invention, the object of the present inventioncan be attained by providing a simple hydraulic circuit device to thebranch circuit.

The present invention according to the modification to the sixthembodiment provides a power setting method for a variable power engine,said variable power engine comprising: a turbocharger for supplying acharged pressure to an engine; a fuel injector with a boost compensatorfor adjusting a fuel oil consumption in accordance with said chargedpressure of said turbocharger; and a transmission for changing outputrevolutions of said engine to a plurality of speed stages, wherein saidcharged pressure of said turbo charger is supplied to said boostcompensator as predetermined pressures of at least two stages equal toor lower than said charged pressure of said turbocharger and higher thanan atmospheric pressure to set an engine power in accordance with aspeed stage position of said transmission.

According to the present invention, a pressure supplied to the boostcompensator can be readily set to two or more stages between the chargedpressure and the atmospheric pressure, thereby easily obtaining adesired engine power.

In the power setting method of the variable power engine according tothe modification of the sixth embodiment, the present invention providesa power setting method, wherein a pressure equal to said chargedpressure of said turbocharger is supplied to said boost compensator toset an engine power when a speed stage position of said transmission isa high speed stage position.

According to the present invention, a light work operation at a highspeed can be assuredly carried out automatically with a high power.

In a vehicle with a variable power engine, a still further object of thepresent invention is to provide a vehicle with a variable power enginewhich can satisfactorily demonstrate performances of the variable powerengine and has the excellent operability and a power controlling methodthereof.

In order to attain the above object, the present invention provides to avariable power engine a transmission with a lockup mechanism and anengine power controller for controlling power of the variable powerengine by a speed stage position signal of the transmission and a lockupenabled/disabled signal of the lockup mechanism.

In addition to the above structure, the present invention furtherprovides a wheel driving state switching mechanism for changing over awheel to be driven among wheels of a vehicle, and a wheel driving statedetection mechanism for detecting the switching state of the wheeldriving stage switching mechanism to output a wheel driving statedetection signal to an engine power controller, thereby furtherachieving the above object.

Specifically, the present invention as disclosed in the eighthembodiment provides a vehicle with a variable power engine comprising:an engine; an engine power switching device for changing a power of saidengine to a plurality of stages; a torque converter with a lockupmechanism coupled with an output side of said engine; a multi-stagespeed change transmission couple to an output side of said torqueconverter; a lockup detection mechanism for detecting on/off of anoperation of said lockup mechanism to output a lockup enabled/disabledsignal; a speed stage position detection mechanism for detecting a speedstage position of said transmission to output a speed stage positionsignal; and an engine power controller to which said speed stageposition signal from said speed stage position detection mechanism andsaid lockup enabled/disabled signal from said lockup detection mechanismare inputted and which outputs an engine power switching signal to saidengine power switching device.

According to the present invention, the torque converter enables thefurther smooth driving; the lockup mechanism can further improve thedriving efficiency at a high speed; and an appropriate engine power canbe obtained in accordance with a speed stage position of thetransmission and the on/off of the lockup of the torque converter.

In the vehicle with the variable power engine according to the eightembodiment, the present invention provides a vehicle with a variablepower engine, wherein a turbocharger for supplying a charged pressure tosaid engine is connected to said engine, and said engine power switchingdevice includes: a fuel injector having a boost compensator foradjusting a fuel oil consumption to said engine in accordance with saidcharged pressure of said turbocharger; and a pressure state switchingmechanism for switching a pressure to be supplied to said boostcompensator to predetermined pressures of at least two stages.

According to the present invention, a power of the variable power enginecan be readily switched by utilizing the charged pressure of theturbocharger.

In the vehicle with the variable power engine according to the eightembodiment, the present invention provides a vehicle with a variablepower engine, wherein a wheel driving state switching mechanism forswitching which wheel among wheels of said vehicle is driven is coupledto said transmission, and said wheel driving state switching mechanismis provided with a wheel driving state detection mechanism for detectinga switching state of said wheel driving state switching mechanism tooutput a wheel driving state detection signal to said engine powercontroller.

According to the present invention, when the wheel driving state can beswitched, the engine power can be optimally controlled in accordance tothis wheel driving state.

The present invention as disclosed in the eighth embodiment provides apower control method of a vehicle with a variable power engine, saidvehicle comprising: an engine; an engine power switching device forchanging power of said engine to a plurality of stages; a torqueconverter with a lockup mechanism coupled to an output side of saidengine; a multi-stage speed change transmission coupled to an outputside of said torque converter; a lockup detection mechanism fordetecting on/off of an operation of said lockup mechanism to output alockup enabled/disabled signal; a speed stage position detectionmechanism for detecting a speed stage position of said transmission tooutput a speed stage position signal; and engine power controller towhich said speed stage position signal from said speed stage positiondetection mechanism and a lockup enabled/disabled signal from saidlockup detection mechanism are inputted and which outputs an enginepower switching signal to said engine power switching device, whereinsaid engine power controller controls said engine power switching devicein such a manner that engine power is obtained on a high power side(high power state) when said speed stage position signal from said speedstage position detection mechanism indicates a higher speed stageposition than a predetermined speed stage and controls said engineoutput switching device in such a manner that said engine power isobtained on said high power side at a higher speed stage position thansaid speed stage position switched to said high power side when saidlockup enabled/disabled signal from said lockup detection mechanism is alockup enabled signal.

According to the present invention, the power of the variable powerengine can be optimally controlled by using the speed stage position ofthe transmission and on/off of the lockup of the torque converter.

The present invention as disclosed in the eighth embodiment provides apower control method of a vehicle with a variable power engine, saidvehicle comprising: an engine; an engine power switching device forchanging power of said engine to a plurality of stages; a torqueconverter with a lockup mechanism coupled to an output side of saidengine; a multi-stage speed change transmission coupled to said outputside of said torque converter; a lockup detection mechanism fordetecting on/off of an operation of said lockup mechanism to output alockup enabled/disabled signal; a speed stage position detectionmechanism for detecting a speed stage position of said transmission tooutput a speed stage position signal; a wheel driving state switchingmechanism which is coupled to an output side of said transmission andswitches which wheel among wheels of said vehicle is driven; a wheeldriving state detection mechanism for detecting a switching state ofsaid wheel driving state switching mechanism to output a wheel drivingstate detection signal; and engine power controller to which said wheeldriving state detection signal from said wheel driving state detectionmechanism, said speed stage position signal from said speed stageposition detection mechanism and said lockup enabled/disabled signalfrom said lockup detection mechanism are inputted and which outputs anengine power switching signal to said engine power switching device,wherein said engine power controller controls said engine powerswitching device in such a manner that engine power is obtained on ahigh power side when said speed stage position signal from said speedstage position detection mechanism indicates a higher speed stageposition than a predetermined speed stage, controls said engine powerswitching device in such a manner that said engine power is obtained onsaid high power side at a higher speed stage position than said speedstage position switched to said high power side when said lockupenabled/disabled signal from said lockup detection mechanism is a lockupenabled signal, and controls said engine power switching device so as toswitch to said high power state at a lower speed stage position than aspeed stage position switched to said high power side before a number ofdriven wheels is increased when said wheel driving state switchingmechanism performs switching in a direction for increasing a number ofdriven wheels.

According to the present invention, the power of the variable powerengine can be optimally controlled by using a speed stage position ofthe transmission, on/off of the lockup of the torque converter and anumber of driven wheels among wheels.

BRIEF DESCRIPTION OF THE DRAWING

FIG. 1 is a block diagram showing a schematic structure of a firstembodiment according to the present invention;

FIG. 2 is a fluid pressure circuit diagram used in the embodiment ofFIG. 1;

FIG. 3 is a graph showing the relationship between an inching pedalangle ratio and a transmission speed ratio in the embodiment of FIG. 1;

FIG. 4 is a flow diagram showing an action of the embodiment of FIG. 1;

FIG. 5 is a graph showing the relationship of an electric currentcorrection value relative to a deviation between a target speed ratioand an actual speed ratio in the embodiment of FIG. 1;

FIG. 6 is a cross-sectional view showing a specific structural exampleof an electronic control regulating valve used in the present invention;

FIG. 7 is a time chart showing the operation of the electronic controlregulating valve of FIG. 6;

FIG. 8 is a drawing for explaining the operation of the electroniccontrol regulating valve of FIG. 6;

FIG. 9 is a drawing for explaining the operation of the electroniccontrol regulating valve of FIG. 6;

FIG. 10 is a drawing for explaining the operation of the electroniccontrol regulating valve of FIG. 6;

FIG. 11 is a drawing for explaining the operation of the electroniccontrol regulating valve of FIG. 6;

FIG. 12 is a drawing for explaining the operation of the electroniccontrol regulating valve of FIG. 6;

FIG. 13 is a drawing for explaining the operation of the electroniccontrol regulating valve of FIG. 6;

FIG. 14 is a block diagram showing a schematic structure of oneembodiment according to the present invention;

FIG. 15 is a graph showing the relationship between a supply current tothe electronic control regulating valve for controlling a clutchmechanism and a clutch mechanism hydraulic oil pressure in theembodiment of FIG. 14;

FIG. 16 is a flow diagram showing an action in the embodiment of FIG.14;

FIG. 17 is a graph showing the relationship between a vehicle speed anda corrected current value in the embodiment of FIG. 14;

FIG. 18 is a graph showing the relationship between the vehicle speedand a clutch mechanism supply oil pressure in another embodimentaccording to the present invention;

FIG. 19 is a graph showing the relationship between the vehicle speedand a driving side clutch mechanism supply oil pressure in still anotherembodiment according to the present invention;

FIG. 20 is a graph showing the relationship between a supply hydraulicpressure to a clutch mechanism on a side opposed to the driving side ineach fixed cycle and the vehicle speed in the embodiment of FIG. 19;

FIG. 21 is a graph showing the relationship between a supply hydraulicpressure supplied to the clutch mechanism on a side opposed to thedriving side in accordance with a deviation (VA-VL) and the vehiclespeed in yet another embodiment according to the present invention;

FIG. 22 is a block diagram showing a schematic structure of the firstembodiment according to the present invention;

FIG. 23 is a graph for explaining a power characteristic in a sixthembodiment;

FIG. 24 is a block diagram showing another embodiment of switching meansin the sixth embodiment;

FIG. 25 is a block diagram showing still another embodiment of theswitching means in the sixth embodiment;

FIG. 26 is a block diagram showing a schematic structure of a seventhembodiment according to the present invention;

FIG. 27 is a block diagram showing another embodiment of a pressurestate switching mechanism in the seventh embodiment;

FIG. 28 is a block diagram showing still another embodiment of thepressure state switching mechanism in the seventh embodiment;

FIG. 29 is a block diagram showing a schematic structure of oneembodiment according to the present invention;

FIG. 30 is a graph showing the relationship between a vehicle speed (V:horizontal axis) and a rim pull (F: horizontal axis) in a vehicle with avariable power engine such as this embodiment; and

FIG. 31 is a graph showing an example of a vehicle performance in ageneral vehicle with a variable power engine.

DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENT(S)

(First Embodiment)

A first embodiment according to the present invention will now bedescribed with reference to the accompanying drawings.

FIG. 1 shows a schematic structure of a primary part of a workingvehicle according to this embodiment.

As similar to a general engine, to an engine 10 are coupled anon-illustrated induction pipe, an exhaust pipe, a fuel injector andothers, and a torque converter with a lockup mechanism 70 is connectedto an output shaft of the engine 10.

The torque converter with the lockup mechanism 70 has the same structureas a general commercial product, and to an output side of the torqueconverter 70 is connected an electronically controlled multi-stage speedchange transmission 30 which contains a gear with a clutch forconverting revolutions of the engine 10 to a multi-stage velocity ratioto be transmitted rearwards. This enables the power of the engine 10 tobe transmitted to the transmission 30 through the torque converter 70.

The transmission 30 changes output revolutions of the engine 10 tomultiple speed stages, for example, eight stages from a forward firstspeed to an eighth speed, four stages from a reverse first speed to afourth speed, namely, a total of 12 speeds to be transmitted to runningmeans 85 such as wheels. The transmission 30 includes a plurality ofdirection switching clutch mechanisms 35 (35A, 35B and 35C) of threestages, i.e., a forward low speed (FL), a forward high speed (FH) andreverse (R) in this embodiment; a plurality of speed switching clutchmechanisms 36 (36A, 36B, 36C and 36D) of four stages, i.e., a firstspeed to a fourth speed in this embodiment; and a gear trainappropriately provided between these clutch mechanisms 35 and 36. Thestructure of the transmission 30 is equal to that of a generalcommercial product.

With this configuration, connecting any of the respective directionswitching clutch mechanisms 35 to any of the speed switching clutchmechanisms 36 can obtain the speed stage position of the forward eightstages and the rear four stages described above.

Specifically, when the forward low speed clutch mechanism FL (35A) ofthe direction switching clutch mechanisms 35 is coupled with any of thefirst to fourth stages (36A, 36B, 36C and 36D) of the speed switchingclutch mechanisms 36, the speed stage positions of the forward first,third, fifth and seventh speeds can be obtained. Further, when theforward high speed clutch mechanism FH (35B) is coupled with any of thefirst to fourth stages of the speed switching clutch mechanisms 36, thespeed stage positions of the forward second, fourth, sixth and eighthspeeds can be obtained. Furthermore, coupling the rear clutch mechanismR (35C) with any of the first to fourth stages of the speed switchingclutch mechanisms 36 can obtain the speed stage positions of the rearfirst, second, third and fourth speeds.

The transmission 30 includes first to third shafts 37, 38 and 39. Aninput side revolution detection mechanism 105 for detecting a number ofrevolutions of the first shaft 37 to output as an input side revolutionsignal N1 to a transmission controller 10 is provided in close vicinityto the first shaft 37 on the uppermost stream side, i.e., on the engine10 side in these shafts. Additionally, an output side revolutiondetection mechanism 106 for detecting a number of revolutions of thesecond shaft 38 to output as an output side revolution signal N2 to thecontroller 100 is provided in close vicinity to the second shaft 38disposed between the direction switching clutch mechanism 35 and thespeed switching clutch mechanism 36. For example, magnetic, optical, orany other type of revolution sensors are used for the revolutiondetection mechanisms 105 and 106.

It is to be noted that an output side revolution detection mechanism106A for detecting a number of revolutions of the third shaft 39 on alowermost stream side, i.e., on the side of running means 35 such aswheels or a tracklayer may substitute for the output side revolutiondetection mechanism 106.

The speed stage position of the transmission 30 can be detected by aspeed stage position detection mechanism (speed stage position signalgeneration mechanism) 31. The speed stage position detection mechanism31 is a mechanism for detecting which speed stage the transmission 30 isselected to, for example, detecting a position of a shift lever 111provided to a driver seat 110 and the like of a working vehicle by usinga detector 112 and outputting a speed stage position signal (TS) to thecontroller 100.

To the driver seat 110 is provided a predetermined clutch mechanism 30,i.e., an inching pedal (clutch pedal) 114 operated for generating a slipin the direction switching clutch mechanism 35 in this embodiment. Anangle sensor 116 constituted by a potentiometer and the like is coupledwith the inching pedal 114 through a link mechanism 115. The linkmechanism 115 and the angle sensor 116 are used to detect a steppingangle of the inching pedal 114, and a pedal angle detection mechanism117 is constituted for outputting a pedal angle signal α to thecontroller 100.

Further, to the driver seat 110 is provided a characteristic changemechanism 118 by which a flexibility of connection (slip condition) ofthe clutch mechanisms 35 by the operation of the inching pedal 114 canbe selected in accordance with an operator's preference. Thischaracteristic change mechanism 118 is constituted by a dial, achangeover switch and others and outputs a characteristic change signalOP to the controller 100.

To the controller 100 are applied a speed stage position signal TS fromthe speed stage position detection mechanism 31, an input siderevolution signal N1 from the input side revolution detection mechanism105, an output side revolution signal N2 from the output side revolutiondetection mechanism 106, a pedal angle signal α from the pedal angledetection mechanism 117 and a characteristic change signal OP from thecharacteristic change mechanism 118, as described above. The controller100 has: an actual speed ratio calculating function 101 for calculatingan actual speed ratio (an actual speed ratio or an actual slip ratio) SAin the transmission 30 based on a ratio of the input side revolutionsignal N1 and the output side revolution signal N2 among the abovesignals; a target speed ratio calculating function 102 for calculating atarget speed ratio SO in the transmission 30 based on the pedal anglesignal α; and a control signal transmitting function 103 for outputtinga connection force control signal CS to a later-described connectionforce control mechanism 120 based on a difference between the targetspeed ratio SO and the actual speed ratio SA (SO-SA) and the speed stageposition signal TS in such a manner that the actual speed ratio SAbecomes a predetermined value.

The controller 100 includes a CPU, a ROM, a RAM and others, andfunctions as a kind of computer so that it can demonstrate theabove-described respective functions and any other function.

The connection force control mechanism 120 is constituted by a pluralityof electronic control regulating valves (ECMV=Electronic controlregulating valve) 121, and each electronic control regulating valve 121is connected to a predetermined clutch mechanism 35 or 36 of themulti-stage speed change transmission 30.

FIG. 2 shows a hydraulic (fluid pressure) circuit including the torqueconverter 70, the transmission 30 and the connection force controlmechanism 120. In this drawing, a plurality of electronic controlregulating valves 121 having the same structure are partiallyspecifically shown, and the remaining part of the same is schematicallyillustrated.

In FIG. 2, a number of the electronic control regulating valves 121 ofthe connection force control mechanism 120 corresponds to a number ofclutch mechanisms of the transmission 30, and the number of the valvesis seven in this embodiment. Here, when the seven electronic controlregulating valves 121 are differentiated from each other, referencecharacters A to G are added to reference numeral 121 of the electroniccontrol valves.

Further, when the direction switching clutch mechanisms 35 and the speedswitching clutch mechanisms of the transmission 30 are differentiatedfrom each other, reference characters A to C are added to referencenumeral 35 of the direction switching clutch mechanisms 35 and referencesymbols A to D are added to reference numeral 36 of the speed switchingclutch mechanisms 36.

In FIG. 2, a strainer 132 is inserted in a tank 131 in which oil as afluid is filled; and a pump 133 is connected to the strainer 132; and apressure oil having a predetermined pressure and a flow rate is suppliedto a main pipe line 134 by the pump 133. An oil filter 135 is providedto the main pipe line 134, and its end is branched off in two.

A hydraulic equipment for the lockup mechanism of the torque converter70 is connected to one branched-off pipe line 136, and the connectionforce control mechanism 120 for controlling the transmission 30 isconnected to the other pipe line 137 through the oil filter 138.

The hydraulic equipment for the lockup mechanism connected to the pipeline 136 includes: a lockup valve 141 consisting of a pilot type 4-port2-position selector valve; a lockup clutch mechanism 142 connected tothe downstream side of the lockup valve 141; and a lockup solenoid valve143 consisting of a three-port two-position selector valve whichsupplies a pilot pressure to the lockup valve 141 and issolenoid-operated.

Here, when no electrical signal is inputted to the solenoid of thelockup solenoid valve 143, the lockup solenoid valve 143 is at anillustrated position by a spring force, and the pilot pressure isapplied to the lockup valve 141 through the lockup solenoid valve 143 sothat the lockup valve 141 is situated at the illustrated positionagainst the spring force, namely it blocks the pipe line 136. Therefore,the hydraulic pressure of the pipe line 136 is not transmitted to thelockup clutch mechanism 142, and the hydraulic pressure in the lockupclutch mechanism 142 flows out to the tank 131 through the lockup valve141, leading to no lockup state of the torque converter 70.

On the other hand, when the electrical signal is inputted to thesolenoid of the lockup solenoid valve 143, the lockup solenoid valve 143is moved to the left side in the drawing against the spring force. Thiscauses the pilot pressure acting on the lockup valve 141 to flow out tothe tank 131 through the lockup solenoid valve 143, and the lockup valve141 is hence moved to the upper side in the drawing by the spring force.Consequently, the hydraulic pressure of the pipe line 136 is supplied tothe lockup clutch mechanism 142 through the lockup valve 141, enablinglockup of the torque converter 70.

An equipment lubricating pipe line 139 is branched off from the middleof the pipe line 137 connected to the connection force control mechanism120 side, and this pipe line 139 is connected to a torque converterlubricating line 145 through a main relief valve 144. The pipe line 139is further connected to a transmission lubricating line 147 through theoil cooler 146 and opened in the tank 131. Moreover, a torque converterrelief valve 148 is provided between the main relief valve 144 of thepipe line 139 and the toque converter lubricating line 145, and atransmission lubricating relief valve 149 is provided between the oilcooler 146 of the pipe line 139 and the transmission lubricating line147.

The torque converter lubricating line 145 is a flow path which isprovided in a body of the torque converter 70 and lubricates the torqueconverter 70 with oil led in the torque converter 70. In addition, thetransmission lubricating line 147 is a flow path which is provided in abody of the transmission 30 and lubricates the transmission 30 with oilled in the transmission 30.

The electronic control regulating valve 121 includes: a pressure controlvalve (proportional control valve) 123 consisting of a four-porttwo-position selector valve having a proportional solenoid 122; and aflow rate detection valve 124 consisting of a pilot type three-porttwo-position selector valve. The pressure control valve 123 converts anoil pressure into another oil pressure responsive to an intensity of acurrent valve of the connection force control signal CS which istransmitted from the controller 100 and received by the proportionalsolenoid 122.

The flow rate detection valve 124 operates by a hydraulic pressuresignal (trigger) based on a pilot hydraulic pressure from the pressurecontrol valve 123 and has the following three functions.

1) The valve is opened until the oil is filled in the directionswitching clutch mechanism 35 and the speed switching clutch mechanism36 of the transmission 30 so that oil filling time of the directionswitching clutch mechanisms 35 and the speed switching clutch mechanisms36 is shortened.

2) The valve is closed at the same time when the oil is filled in thedirection switching clutch mechanism 35 and the speed switching clutchmechanism 36, and a signal (fill signal) is outputted to the controller100 to inform of completion of filling.

3) When the hydraulic pressure is being applied to the directionswitching clutch mechanism 35 and the speed switching clutch mechanism36, the fill signal is outputted to the controller 100 to inform ofpresence/absence of the hydraulic pressure.

In FIG. 2, the respective electronic control regulating valves 121constituting the connection force control mechanism 120 are connected tothe respective clutch mechanisms 35 and 36 of the transmission 30.

Specifically, the electronic control regulating valve 121A is connectedto the forward low speed clutch mechanism FL (35A) of the directionswitching clutch mechanism 35; the electronic control regulating valve121B, to the forward high speed clutch mechanism FH (35B) of thedirection switching clutch mechanism 35; the electronic controlregulating valve 121C, to a rear clutch mechanism R (35C) of thedirection switching clutch mechanism 35; the electronic controlregulating valve 121D, to the first stage clutch mechanism 36A of thespeed switching clutch mechanism 36; the electronic control regulatingvalve 121E, to the second stage clutch mechanism 36B of the speedswitching clutch mechanism 36; the electronic control regulating valve121F, to the third stage clutch mechanism 36C of the speed switchingclutch mechanism 36; and the electronic control regulating valve 121G,to the fourth clutch mechanism 36D of the speed switching clutchmechanism 36.

Although the structure of the respective direction switching clutchmechanism 35 and speed switching clutch mechanisms 36 in which a springis used as a drive source for separating a clutch is illustrated, thepresent invention is not necessarily restricted thereto, and it may be arotating clutch type structure which operates by a pilot pressure asindicated by a virtual line (alternate long and two short dashes line)33 at a position of the forward low speed clutch mechanism FL (35A) inFIG. 2.

(Results of First Embodiment)

Results of this embodiment configured as described above will now bedescribed with reference to graphs of FIGS. 3 and 5 and a flow chart ofFIG. 4.

Referring to FIG. 2, at the start of the operation using the workingvehicle, when the engine 10 is driven and the forward first speed isselected by the shift lever 111, the speed stage position detectionsignal TS is sent from the speed stage position detection mechanism 31to the controller 100, and the connection force control signal CS issupplied from the controller 100 to the connection force controlmechanism 120. The connection force control signal CS causes theelectronic control regulating valve 121 of the connection force controlmechanism 120 to be operated in such a manner that the forward low speedclutch mechanism FL (35A) of the direction switching clutch mechanisms35 is connected to the first stage clutch mechanism 36A of the speedswitching clutch mechanism 36 in the transmission, thereby starting theworking vehicle.

At this time, since the lockup signal is not inputted to the lockupsolenoid valve 143, the torque converter 70 is not locked up and therebyfunctions as a normal torque converter.

Incidentally, the lockup operation of the torque converter 70 is notconcerned in the present invention, the explanation thereof is omitted,but the lockup operation is carried out during the high speed runningand the like as similar to the general cases.

A signal of the electronic control regulating valve 121 to which theconnection force control signal CS has been transmitted from thecontroller 100 is inputted to the proportional solenoid 122 shown inFIG. 2, and the pressure control valve 123 moves to the left side inresponse to a current amount of the signal against the spring force. Asa result, the hydraulic pressure supplied from the pump 133 through themain pipe line 134 and the pipe line 137 flows into the flow ratedetection valve 124 through the pressure control valve 123 and issupplied to the forward low speed clutch mechanism FL (35A) and thefirst stage clutch mechanism 36A of the speed switching clutch mechanism36.

Thereafter, the sequential shift up is performed by the shift lever 111so that the shift lever 111 is set to the speed stage position matchedwith a required work. For example, when working at the forward thirdspeed, the forward low speed clutch mechanism FL (35A) of the directionswitching clutch mechanisms 35 is connected to the second stage clutchmechanism 36B of the speed switching clutch mechanisms 36.

In this state, when reduction in the vehicle speed is desired whilemaintaining the power for the work, the inching pedal 114 is stepped.This can cause the pedal angle signal α from the angle sensor 116 of thepedal angle detection mechanism 117 to be outputted to the controller100, and the connection force control signal CS from the controller 100lowers the current value outputted to the electronic control regulatingvalve 121 of the connection force control mechanism 120. Reduction inthe current value involves decrease in the connection force of thedirection switching clutch mechanism 35A to cause a slip in thedirection switching clutch mechanism 35A, thus lowering the vehiclespeed.

FIG. 3 shows the relationship between an angle of the inching pedal 114described above and a speed ratio S of the input side and the outputside in the transmission 30.

In FIG. 3, a horizontal axis represents an angle ratio (%) of theinching pedal 114 and a vertical axis represents a speed ratio (%) ofthe input side and the output side of the transmission 30.

Here, 0% of the inching pedal angle ratio corresponds to the state wherethe inching pedal is not stepped at all; 100% of the same, the statewhere the inching pedal is stepped completely; 10% of the same, thestate where a foot is lightly put on the inching pedal 114 and there isno will of stepping on it; and 90% of the same, the state where theinching pedal is completely stepped while taking into account machineirregularities and the like.

Further, 0% of the speed ratio S corresponds to the complete slippage,i.e., no connection is established; and 100%, the complete connection isestablished with no slip.

In FIG. 3, when stepping of the inching pedal 114 is less than 10%, alarge current is supplied to the proportional solenoid 122 of theelectronic control regulating valve 121 so that the speed ratio Sbecomes 100%, and when stepping is not less than 90%, no current issupplied so that the speed ratio S becomes 0%. In the intermediateposition, as indicated by a solid line M, the relationship between anangle of the inching pedal 114 and the speed ratio S is designed tolinearly vary and, for example, 50% of the speed ratio (slip ratio) Scan be obtained with the 50% stepping of the inching pedal 114.

Here, the control target causing slip is determined to be the directionswitching clutch mechanism 35 as described above.

Further, although an angle of the inching pedal 114 is detected by theangle sensor 116, if the angle sensor 116 is constituted by apotentiometer, the detected angle is outputted as a voltagecorresponding thereto to the controller 100. The characteristic obtainedby the potentiometer is represented by a notation in parentheses of thehorizontal direction in FIG. 3.

When operating the inching pedal 114, the stepping angle and the senserelative to the slip ratio are different depending on the operators.Therefore, in the present embodiment, the characteristics indicated byalternate long and two short dashes lines P, Q and R are presented inaddition to the characteristic indicated by the solid line M. Here, whenthe large slip with a small stepping quantity is desired, thecharacteristic of the long and two short dashes line R may be selected.Also, when increase in slip is desired with a stepping angle which islarge in some measure, and when the intermediate characteristic isdesired, P or Q may be selected.

Switching of these characteristics can be attained by manipulating adial and the like of a characteristic change mechanism 118 by anoperator, and the controller 100 performs control responsive to themanipulation state of the characteristic change mechanism 118.

Although the above has described the cases where slippage is caused inthe direction switching clutch mechanism 35 by stepping on the inchingpedal 114 in running of the working vehicle, i.e., during the workingstate, the similar operation is carried out when the machine is stoppedand the inching pedal 114 is gradually released from the completelystepped state (not less than 90%) of the inching pedal 114 to start themachine.

In such a case, although the matching of the feeling (sense) with startof the machine enabled by a given level of pedal releasing may differdepending on the operators, this can be dealt by operating theabove-described characteristic change mechanism 118.

Further, the slip rate relative to the stepping angle varies inaccordance with the speed stage position.

That is, in general, it is determined that the speed ratio (speeddecreasing ratio) of the input side and the output side in the forwardfirst speed is, for example, approximately 10, whereas it isapproximately two in case of the forward fifth speed. Also, it isapproximately 0.8 in the forward eighth speed (speed increasing ratio).

In terms of the driving force (torque) in the working vehicle, thelarger torque can be demonstrated as the speed stage position becomeslower. On the other hand, in order to drive the working vehicle, thedriving force above a give value is required, and the necessary drivingforce for moving the working vehicle is substantially fixed irrespectiveof the speed stage position.

Therefore, in the low speed stage position such as the forward firstspeed by which the large torque can be demonstrated, starting is enabledeven if the considerably large slippage is generated in the directionswitching clutch mechanism 35, i.e., even if the inching pedal 114 isslightly released from the maximum stepped state (the state in which theinching pedal angle ratio is close to 90%). Whilst, when trying to startfrom the medium speed stage position such as the forward third or fourthspeed, since the original torque (driving force) on the output side ofthe transmission 30 is small, starting is not possible unless theinching pedal 114 is largely released from the maximum stepping state.

In such a case, it is also preferable for the operator to enable thesimilar operation, i.e., starting or deceleration with the similarstepping angle of the inching pedal 114.

As to the control for enabling starting or acquirement of the runningspeed with the same sense of the pedal operation in various workingsituations described above, it is preferable for the operator to enablestarting the vehicle on a grade or acquirement of the running speedduring the operation with the same sense of the pedal operation as wellas the above described cases. That is, in starting of the vehicle, whengradually releasing the fully pressed down inching pedal 114, startingis not enabled unless a releasing amount is large on the up grade.Meanwhile, on the down grade, a small amount of releasing enablesstarting the vehicle, and the sense relative to the operation differsbetween the up grade and the down grade. There is a request for enablingthese operations with the same sense of the operation of the inchingpedal 114.

Therefore, in the present embodiment, monitoring by the controller 100the input side revolution signal N1 and the output side revolutionsignal N2 of the transmission 30 as well as the speed stage positionsignal TS from the shift lever 111 of the speed stage position detectionmechanism 31 or the pedal angle detection signal α from the pedal angledetection mechanism 117 of the inching pedal 114 causes the controller100 to control so that the same output revolution number can be obtainedwith respect to the same angle of the inching pedal 114 irrespective ofthe speed stage position.

A flowchart of FIG. 4 shows the procedure of this control.

In FIG. 4, when the control starts, the inching pedal angle signal α,the speed stage position signal TS, the characteristic change signal OP,and the input side revolution signal N1 and the output side revolutionsignal N2 of the transmission 30 are read at fixed time intervals to beupdated in the controller 100 in the step 201.

In the step 202, a target speed ratio of the input side and the outputside, i.e., a target speed ratio SO is calculated based on the inchingpedal angle signal α, the speed stage position signal TS and thecharacteristic change signal OP, and an actual speed ratio SA isobtained based on the input side revolution signal N1 and the outputside revolution signal N2.

Subsequently, in the step 203, a deviation (SO-SA) between the targetspeed ratio SO and the actual speed ratio SA is calculated and ajudgment is made upon whether this deviation (SO-SA) is within apredetermined deviation (dead zone).

If it is within the predetermined deviation, the supply current to theproportional solenoid 122 in the pressure control valve 123 forcontrolling the direction switching clutch mechanism 35 which is undercontrol is maintained without changing the current indicated value inthe step 204, and the control returns to the step 201 to repeat the step202 and the following steps.

On the other hand, if the deviation (SO-SA) is out of the predetermineddeviation, a judgement is made upon whether this value is above thepredetermined deviation (+) in the step 205.

If (+), in the step 206, the supply current to the proportional solenoid122 in the pressure control valve 123 for controlling the directionswitching clutch mechanism 35 which is under control is specified to thecurrent indicated value by adding a correction value of a predeterminedcorrection map, and the control returns to the step 201 to repeat thestep 202 and the following steps.

If it is not (+), i.e., (−), in the step 207, the supply current to theproportional solenoid 122 in the pressure control valve 123 forcontrolling the direction switching clutch mechanism 35 which is undercontrol is specified to the current indicated value by subtracting acorrection value of the predetermined correction map, and the controlreturns to the step 201 to repeat the step 202 and the following steps.

FIG. 5 shows the relationship between the above-described deviation andthe current correction value.

In FIG. 5, the horizontal axis presents the deviation and the verticalaxis shows the current correction value. If the deviation (SO-SA) is notvery large and is within a predetermined range, the current correctionis not performed. That is because control for correcting a smalldeviation causes hunching to occur and the control may become unstable.Namely, a predetermined deviation functions as a so-called dead zone.

When the deviation increases beyond the dead zone, the currentcorrection value is linearly increased. Meanwhile, when the deviationdecreases beyond the dead zone, the current correction value is linearlyreduced. That is, since increase in the deviation means that adifference in number of revolutions between a target value and an actualmeasured value and the slip is large, the value of the current suppliedto the proportional solenoid 122 of the pressure control valve 123 isincrease to give rise to an oil amount supplied to the directionswitching clutch mechanism 35. Further, the connection force of theclutch mechanisms 35 is enhanced to reduce the slip. If the deviation isdecreased, the reverse operation is applied.

Although the speed stage position signal TS is taken into considerationin advance to be set when calculating the target speed ratio SO in theflowchart of FIG. 4, the setting of the target speed ratio SO is notrestricted thereto. For example, the target speed ratio SO may be setbased on only the pedal angle signal a and the target speed ratio SO maybe compared with the actual speed ratio SA. Thereafter, correction maybe carried out in accordance with the speed stage position signal TS soas to control the current value to the electronic control regulatingvalve 121 and the connection force of the clutch mechanism 35.

Further, when starting the vehicle on a down grade and the slop issteep, it can be considered that the running speed of the workingvehicle may exceeds a desired speed due to the slope of the grade evenif supply of the oil to the direction switching clutch mechanism 35 isall interrupted.

In such a case, since the running speed can not be decreased to thatextent even if the inching pedal 114 is fully stepped, there is effectedthe control that the oil is supplied to the reverse direction switchingmechanism 36 to be operated as a brake for the working vehicle in adirection opposed to the used (driven) direction switching clutchmechanism 35 when driving this forward direction switching clutchmechanism 35.

Further, there is carried out the different control that the forward andreverse clutch mechanisms 35 and 36 are simultaneously activated to copewith starting the vehicle on a grade. In such a case, the clutchmechanism on the opposed side always acts as a brake in some smallmeasure, thereby smoothing the control.

(Advantages of the First Embodiment)

According to the above embodiment, the following advantages can beobtained.

(Advantages 1-1) In this embodiment, when controlling the connectionforce of the direction switching clutch mechanism 35 by the inchingpedal 114, since the controller 100 monitors and controls the inchingpedal angle signal α, the speed stage position signal TS, as well as theinput side revolution signal N1 and the output side revolution signal N2of the transmission 30, it is possible to effect the appropriate controlaccording to the actual slip in the transmission 30.

In particular, the fixed running state can be always obtained withrespect to the stepping state of the inching pedal 114 irrespective ofthe load state in the work or the slope state such as an up grade or adown grade.

(Advantage 1-2) Since control of the clutch force in the transmission 30is carried out by the electron control modulation valve 121 connected toeach direction switching clutch mechanism 35, the optimum control ispossible in accordance with each direction switching clutch mechanism35.

Here, since the electron control modulation valve 121 having the similarstructure is connected to each speed switching clutch mechanism 36, eachspeed switching clutch mechanism 36 can be precisely controlled ifnecessary.

(Advantage 1-3) The connection force control mechanism 120 includes: thepressure control valve 123 which receives the connection force controlsignal CS from the controller 110 and converts the hydraulic pressure tothe counterpart responsive to the received signal CS; and the flow ratedetection valve 124 which operates in accordance with the hydraulicpressure signal from the pressure control valve 123, thus enabling theappropriate control according to the direction signal (connection forcecontrol signal CS) with a relatively simple structure.

(Advantage 1-4) Since the clutch mechanism which is controlled by thecontroller 100 and generates a slip in accordance with an amount ofoperation by the inching pedal 114 is restricted to the directionswitching clutch mechanisms 35 as a part of the clutch mechanism of thetransmission 30, cooling means for cooling down the heat generated bythe slip can cope with only the direction switching clutch mechanisms 35whose number is relatively small, thereby reducing the cost of theapparatus.

As a specific example, provision of a pipe arrangement and a controlvalve for increasing a supply amount of oil for cooling (lubricating)the direction switching clutch mechanism 35 which generates a slip cansuffice the cooling means.

(Advantage 1-5) Since a full-speed stage in the transmission 30 isachieved by the combined use of the direction switching clutch mechanism35 and the speed switching clutch mechanism 36, the inching pedal 114can perform the slip control over the all the speed stages if thedirection switching clutch mechanism 35 can generate a slip, thusenabling the necessary and sufficient control with a small number ofcontrol targets.

(Advantage 1-6) Since the controller 100 controls the deviation (SO-SA)in such a manner that the output side revolution N2 of the transmission30 can be a number of revolutions within a predetermined deviation,generating no hunching which leads to the unstable control.

(Advantage 1-7) Since the characteristic change dial is provided as thecharacteristic change mechanism 118 to the driver seat 110 and thecharacteristic change signal OP from the characteristic change mechanism118 is inputted to the controller 100, the content of the connectionforce control signal CS outputted from the controller 100 to theconnection force control mechanism 120 can be changed in accordance withthe work conditions or the operator's preference.

(Specific Example of the Electronic Control Regulating Valve of theFirst Embodiment)

FIG. 6 shows a specific structural example of the electronic controlregulating valve 121 used in this embodiment.

In FIG. 6, the electronic control regulating valve 121 includes a valvebody 153 having two valve holes 151 and 152 and a cover body 154 forblocking one end of the valve body 153. A lower half portion of the body153 forms the pressure control valve 123 and an upper half portion ofthe same forms the flow rate detection valve 124.

The proportional solenoid 122 is provided so as to be opposed to thevalve hole 151 on the opposite side of the cover body 154 of the valvebody 153. The proportional solenoid 122 includes an electromagnet 155and a core 156 which can move along the axial direction in theelectromagnet 155. An end of the core 156, which is the left end in thedrawing, is brought into contact with one end of a pressure controlvalve spool 157 which is housed in the valve hole 151 so as to becapable of sliding. A pressure control valve spring 158 consisting of ahelical compression spring is provided between the other end of thepressure control valve spool 157 and the cover body 154, and the spring158 always gives an impetus to the pressure control valve spool 157 inthe right-hand side direction in the drawing.

The pressure control valve spool 157 has first to third land portions161, 162 and 163 in the middle part thereof and further includes aninternal communication hole 164 which pierces through the central secondland portion 162 and extends to the left end of the same along the axialdirection. A load piston 165 is accommodated in the internalcommunication hole 164 along the axial direction so as to be capable ofsliding, and the left end of the load piston 165 projects from theinternal communication hole 164 to be brought into contact with theinner surface of the cover body 154.

Three communication holes 166, 167 and 168 for communicating the valvehole 151 of the pressure control valve 123 to the valve hole 152 of theflow rate detection valve 124 are provided in the valve body 153. Thecentral communication hole 167 serves as pump ports A and D for thepressure control valve 123 and the flow rate detection valve 124 and isconnected to the pump 133 through the pipe line 137, the pump 133 beingdesigned to suck in the oil from the tank 131 through the main pipe line134.

In the valve hole 151, the intermediate position between the centralsecond land portion 162 and the left third land portion 163 and the leftposition of the left third land portion 163 communicate with the rightend communication hole 166 through a drain communication hole 169, andthe drain communication hole 169 functioning as a drain port E so as tobe capable of discharging the oil to the tank 131 through the pipe line171.

Further, the left communication hole 168 can communicate with or can beshut off from the intermediate position between the central second landportion 162 and the left third land portion 163 or the centralcommunication hole 167 by the operation of the second land portion. Aposition of the communication hole 168 where it crosses the valve hole151 functions as a pressure control valve output port B.

A fill switch 172 is provided so as to be opposed to the valve hole 152at an opposite position of the cover body 154 of the valve body 153. Thefill switch 172 has an actuator 173 which can move along the axialdirection. An end of the actuator 173, which is the left end in thedrawing, is opposed to one end of the flow rate detection valve spool174 which is housed in the valve hole 152 so as to be capable ofsliding, and a fill switch spring 175 consisting of a helicalcompression spring is provided between one end of the flow ratedetection valve spool 174 and a casing for the fill switch 172. On theother hand, a flow rate detection valve return spring 176 consisting ofa compression spring is provided between the other end of the flow ratedetection valve spool 174 and the cover body 154, and the operation ofthese springs 175 and 176 give an impetus to the flow rate detectionspool 174 so as to be balanced at a predetermined position.

The flow rate detection valve spool 174 has three first to third landportions 177, 178 and 179 in the middle part thereof and furtherincludes a plurality of orifices 181 piercing through the left thirdland portion 179 in the axial direction.

Here, the valve hole 152 communicates with the communication hole 166 ata right position of the first land portion 177, with the communicationhole 167 at an intermediate position between the first land portion 177and the second land portion 178 and with the communication hole 168 atan intermediate position between the second land portion 178 and thethird land portion 179, respectively. Further, the left side of thethird land portion 179 serves as a clutch port C and communicates withany clutch mechanism of the direction switching clutch mechanism 35 andthe speed switching clutch mechanism 36 through the pipe line 182.

It is to be noted that the communication hole 167 and the communicationhole 168 can communicate with or can be blocked off from each other bythe second land portion 178.

The operation of the electronic control regulating valve 121 accordingto this embodiment will now be described with reference to FIG. 8 toFIG. 13.

The electronic control regulating valve 121 is controlled by a directioncurrent from the controller 100 to the proportional solenoid 122 of thepressure control valve 123 and an output signal of the fill switch 172.

FIG. 7 shows the relationship between the direction current of theelectronic control regulating valve 121 to the proportional solenoid122, an input pressure of the clutch mechanisms 35 and 36 and an outputsignal of the fill switch 172.

That is, as shown in FIG. 7, a region H before a point in time T1represents before the pressure control valve 123 is selected for speedchange, and the oil of the clutch mechanisms 35 and 36 is drained.

Subsequently, in a region I between points in time T1 and T2, a largecurrent instruction value is supplied as a hydraulic pressure signal tothe proportional solenoid 122 to start filling. At the point in time T2after lapse of a fixed time from the point in time T1 at which fillingis started, the instruction current to the proportional solenoid 122 isonce lowered to a predetermined initial value. Here, the instructioncurrent is caused to be maximum at the start of filling in order torapidly fill the oil to the clutch mechanisms 35 and 36.

Filling with the initial value is continued until a point in time T3(region J), during which temporal filling of oil to the clutchmechanisms 35 and 36 is completed, and hence the hydraulic pressure ofthe clutch mechanisms 35 and 36 is increased, thereby turning on thefill switch 172.

A region K after the point in time T3 is a pressure regulation region Kin which the instruction current to the proportional solenoid 122 isregulated in such a manner that the connection force of the clutchmechanisms 35 and 36 becomes a predetermined value.

A region L obtained by combining the regions I and J represents filling.

Among the electronic control regulating valves 121 constituting theconnection force control mechanism 120, the controller 100 also controlsthe electronic control regulating valve 121 for the speed switchingclutch mechanism 36 which does not relate to a slip based on theoperation of the inching pedal 114, but the explanation of the speedswitching clutch mechanism 36 is omitted.

The specific operation of the electronic control regulating valve 121will now be described with reference to FIGS. 8 to 13.

FIG. 8 shows the state before the speed change where the oil is drainedfrom the direction switching clutch mechanism 35, the statecorresponding to the region H in FIG. 7.

In this state where no current is passed through the electromagnet 155of the proportional solenoid 122, the core 156 of the proportionalsolenoid 122 is returned by the repulsion force of the pressure controlvalve spring 158 through the pressure control valve spool 157.Therefore, the second land portion 162 of the pressure control valvespool 157 moves in the right direction and the communication hole 168communicates with the drain communication hole 169 so that the oil ofthe clutch port C is drained to the tank 131 through the orifice 181provided to the third land portion 179 of the flow rate detection valvespool 174, the communication hole 168, the output port B and the drainport E.

Here, since the hydraulic pressure does not act on the flow ratedetection valve spool 174 of the flow rate detection valve 124, the flowrate detection valve spool 174 is removed from the actuator 173 of thefill switch 172 by the repulsion force of the fill switch spring 175 andstopped at a position balanced with the spool return spring 176.

FIG. 9 shows the state of starting filling where the instruction currentbased on the hydraulic pressure signal is inputted to the electromagnet155 of the proportional solenoid 122, the state corresponding to theregion I of FIG. 7.

When the maximum current is supplied to the proportional solenoid 122 asa current based on the hydraulic pressure signal with no oil beingfilled in the direction switching clutch mechanism 35, the core 156 ofthe proportional solenoid 122 demonstrates full stroke and the pressurecontrol valve spool 157 moves to the left-hand direction. This causesthe second land portion 162 to move to the left-hand side, and the pumpport A and the pressure control valve power port B are opened while theoutput port B and the drain port E are blocked. Accordingly, the oilflows to the clutch port C through the pump port A, the output port Band the orifice 181 of the flow rate detection valve spool 174 so thatthe oil is filled in the direction switching clutch mechanism 35.

At this time, the throttling effect of the orifice 181 generates adifferential pressure between the upstream and the downstream of theorifice 181 of the flow rate detection valve spool 174. Thisdifferential pressure causes the flow rate detection valve spool 174 tomove in the left direction while compressing the flow rate detectionvalve spool return spring 176 as shown in FIG. 10. This state stillcorresponds to the region I in FIG. 7.

The leftward movement of the flow rate detection valve spool 174 causesthe pump port D of the flow rate detection valve 124 to open, and theoil flows from this port toward the communication hole 168. Then, thedifferential pressure between the upstream and the downstream of theorifice 181 of the flow rate detection valve spool 174 becomes large,which further moves the flow rate detection valve spool 174 in the leftdirection.

In this state, when the current value of the proportional solenoid 122is instantaneously lowered to the initial pressure level, filling isterminated and the pressure control valve 123 is reset to the initialpressure as shown in FIG. 11, the state of which corresponds to theregion J in FIG. 7.

That is, in this state, the leftward driving force of the core 156 islowered to return to the right direction on one hand, and the pressureof the pressure control valve output port B of the communication hole168 is led in the internal communication hole 164 on the other hand.

Since this can substantially apply the pump pressure to the load piston165, the pressure control valve spool 157 is again pushed toward theright side, and a small amount of oil leaks from the pressure controlvalve power port B to the drain port E.

However, this amount of leak is small, a large amount of the oil flowsfrom the pump 133 to the direction switching clutch mechanism 35, andthe flow rate detection valve spool 174 is continuously pushed towardthe left side.

In this manner, when the oil is filled in the direction switching clutchmechanism 35, the oil does not flow from the pump port D to the clutchport C.

Subsequently, the state shown in FIG. 12 is obtained, which stillcorresponds to the region J in FIG. 7.

When a flow of the oil to the clutch port is eliminated and thehydraulic pressures on the both sides of the third land portion becomeequal, the flow rate detection valve spool 174 moves to the right sideby the hydraulic pressure since the pressure applied areas on the bothsides of the third land portion 179 in the flow rate detection valvespool 174 differ from each other and the pressure applied area on theleft side is larger. As a result, the pump port D and the clutch port Care closed.

Here, the flow rate detection valve spool 174 is pushed toward the rightside by a difference in area between the right and left sides of thethird land portion 179 and the force of the flow rate detection valvespool return spring 176 while compressing the fill switch spring 175 andbrought into contact with the actuator 173 of the fill switch 172 inorder to inform the controller 100 of completion of filling to thedirection switching clutch mechanism 35. At this time, since the currentvalue at the initial pressure level is given to the proportionalsolenoid 122, the hydraulic pressure by the pressure control valve spool157 is set to the initial pressure.

FIG. 13 shows the state where the pressure is regulated, whichcorresponds to the region K in FIG. 7.

After the initial pressure setting, when a predetermined current iscaused to flow to the proportional solenoid 122, the electromagnet 155of the proportional solenoid 122 generates a force in proportion to thecurrent. A sum of the thrust of the proportional solenoid 122, thethrust obtained by the hydraulic pressure of the clutch port C appliedto the load piston 165 and the repulsion force of the pressure controlvalve spring 158 is balanced to achieve pressure regulation.

On the other hand, since the flow rate detection valve spool 174 iscontinuously pushed toward the right side by a difference in hydraulicpressure applied to the both sides of the third land portion 179, thefill switch 172 continues to output a fill signal to the controller 100.

Thereafter, when the speed stage position is changed by the operation ofthe shift lever 111 to disable the direction switching clutch mechanism35, the current supplied to the proportional solenoid 122 is stopped.Therefore, the pressure control valve spool 157 is moved to the rightside by the spring force of the pressure control valve spring 158 andreset to the state shown in FIG. 8, thereby entering the standby modefor next use.

According to the electron control modulation valve 121 described above,the following advantages can be obtained.

(Advantage 1-8) The pressure control valve 123 and the flow ratedetection valve 124 can be incorporated in the integral valve body 153,thereby providing the small and inexpensive apparatus.

(Advantage 1-7) The pressure control valve spool 157 and the flow ratedetection valve spool 174 are provided in parallel to each other, andopenings of the valve hole 151 and the valve hole 152 on one side areblocked by the cover body 154 while those on the other side are blockedby the proportional solenoid 122 and the fill switch 172. Therefore,assembling can be facilitated by one direction that is the horizontaldirection in the drawing, which leads to the inexpensive production ofthe apparatus.

(Modification of the First Embodiment)

It is to be noted that the present invention is not restricted to theembodiment shown in FIGS. 1 to 5 or FIG. 6, and any modification orimprovement in the scope of the invention can be included in the presentinvention.

For example, the connection force control mechanism 120 is notnecessarily the electronic control regulating valve 121 and a differenttype of control valve may be used. However, use of the electroniccontrol regulating valve 121 has such an advantage as that theconnection force responsive to the stepping angle of the inching pedal114 can be readily obtained with a simple structure.

Here, although the electronic control regulating valve 121 having thestructure shown in FIG. 6 does not have to be used, use of such a valvecan inexpensively provide a small apparatus.

Further, the clutch mechanism whose connection force is controlled inaccordance with the stepping angle of the inching pedal 114 is notrestricted to the direction switching clutch mechanism 35, only thespeed switching clutch mechanism 36 or the both mechanisms may be used.However, the direction switching clutch mechanism 35 has such anadvantage as that all the speed stage positions can be controlled with asmall number of control targets.

Moreover, although a number of revolutions on the output side of thetransmission 30 does not have to be controlled so as to be within apredetermined deviation range, such a control can prevent hunching.

In addition, provision of the characteristic change dial 118 as thecharacteristic change mechanism is not necessary, but it is advantageousin that control is enabled in accordance with the operator's preference.

(Second Embodiment)

A second embodiment according to the present invention will now bedescribed with reference to FIG. 14 and FIG. 2 described above.

The basic structure of this embodiment is similar to that of the firstembodiment, and like reference numerals denote like or correspondingparts to omit the explanation. The following will describe differentparts.

The transmission 30 has three shafts, i.e., first to third shafts 37, 38and 39. The output side revolution detection mechanism 106A, whichdetects a number of revolutions of the third shaft 39 to be outputted asthe output side revolution signal N2 to the transmission controller 100,is provided in close vicinity to the shaft on the lowermost stream sideamong the above-described three shafts, i.e., the third shaft 39 on theside of the running means 85 such as wheels or a tracklayer. Forexample, a magnetic, optical or any other type of revolution sensor isused for the revolution detection mechanism 36A.

Incidentally, in this embodiment, the output side revolution detectionmechanism 106A may be provided in close vicinity to the second shaft 38disposed between the direction switching clutch mechanism 35 and thespeed switching clutch mechanism 36 in order to detect a number ofrevolutions of the second shaft 38 to be outputted as the output siderevolution signal N2 to the controller 100.

The speed stage position of the transmission 30 is detected by the speedstage position detection mechanism (speed stage position signalgeneration mechanism) 31. The speed stage position detection mechanism31 detects which speed stage the transmission 30 is selected to bydetecting, for example, a position of the shift lever 111 provided tothe driver seat 110 of the working vehicle by using the detector 112 andoutputs the speed stage position signal (TS) to the controller 100.

The accelerator pedal 107 operated for increasing revolutions of theengine 10 is provided to the driver seat 110. To the accelerator pedal107 is coupled the angle sensor 109 constituted by a potentiometer andthe like through the link mechanism 108, and the link mechanism 108 andthe angle sensor 109 form the accelerator pedal angle detectionmechanism 113. The accelerator pedal angle detection mechanism 113detects a stepping angle of the accelerator pedal 107 to output anaccelerator pedal angle signal a to the controller 100.

The speed mode setting mechanism 119 is provided to the driver seat 110if necessary. The speed mode setting mechanism 119 can switch thevehicle speed between the normal running mode and the hyper-slow runningmode and includes a dial, a changeover switch and others. It outputs arunning mode signal OM to the controller 100.

As described above, the controller 100 receives the speed stage positionsignal TS from the speed stage position detection mechanism 31, theoutput side revolution signal N2 from the output side revolutiondetection mechanism 106A and the accelerator pedal angle signal a fromthe accelerator pedal angle detection mechanism 113 and, when the speedmode setting mechanism 119 is provided, the controller 100 furtherreceives the running mode signal OM from the speed mode settingmechanism 119.

The controller 100 for receiving these signals includes: a hyper-slowrunning mode judging function 104 for judging whether it is thehyper-slow running mode based on a vehicle speed V calculated from theoutput side revolution signal N2, the accelerator pedal angle signal âand the speed stage position signal TS or the running mode signal OM;and a control signal transmitting function 103 for outputting aconnection force control signal CS to the later-described connectionforce control mechanism 120 so that a number of revolutions on theoutput side of the transmission 30 becomes a predetermined hyper-slowvalue when the hyper-slow running mode is determined by the hyper-slowrunning mode judging function 104.

The controller 100 has a CPU, a ROM, a RAM and the like and functions asa kind of computer so as to demonstrate the above-described respectivefunctions and any other function.

The connection force control mechanism 120 is constituted by a pluralityof electronic control regulating valves (ECMV=Electronic controlregulating valve) 121, and each electronic control regulating valve 121is connected to a predetermined mechanism in the respective clutchmechanisms 35 and 36 of the multi-stage speed change transmission 30.

Since the connection force control mechanism 120 is the same as thatdescribed in the first embodiment, thereby omitting its explanation.

(Results of the Second Embodiment)

The Results of the embodiment having the above structure will bedescribed with reference to graphs of FIGS. 15 and 17 and a flowchart ofFIG. 16.

In FIGS. 14 and 2, at the start of the operation of the working vehicle,when the engine 10 is driven and the forward first speed is selected bythe shift lever 111, the speed stage position signal TS is supplied fromthe speed stage position detection mechanism 31 to the controller 100,and the connection force control signal CS is sent from the controller100 to the connection force control mechanism 120. The electroniccontrol regulating valve 121 of the connection force control mechanism120 is activated by the connection force control signal CS in such amanner that the forward low speed clutch mechanism FL (35A) of thedirection switching clutch mechanism 35 is connected to the first stageclutch mechanism 36A of the speed switching clutch mechanism 36 in thetransmission 30, thus starting the working vehicle.

At this time, since the lockup signal is not inputted to the lockupsolenoid valve 143, the torque converter 70 is not locked up and therebyfunctions as a normal torque converter. The lockup operation of thetorque converter 70 does not relate to the present invention, and itsexplanation is omitted, but the lockup operation is carried out duringthe high speed running and the like as similar to the general case.

The electronic control regulating valve 121, which has received theconnection force control signal CS from the controller 100, inputs itssignal to the proportional solenoid 122 shown in FIG. 2, and thepressure control valve 123 moves to the left side against the springforce in accordance with a current amount of that signal. Consequently,the hydraulic pressure supplied from the pump 133 through the main pipeline 134 and the pipe line 137 flows in the flow rate detection valve124 via the pressure control valve 123 to be further supplied to theforward low speed clutch mechanism FL (35A) and the first stage clutchmechanism 36A of the speed switching clutch mechanism 36.

Thereafter, the shift up is sequentially conducted by the shift lever111, and the shift lever 111 is set to a speed stage position suitablefor a required operation. For example, in case of operating at theforward third speed, the forward low speed clutch mechanism FL (35A) ofthe direction switching clutch mechanism 35 is connected to the secondstage clutch mechanism 36B of the speed switching clutch mechanism 36.

The shift up is then appropriately carried out to perform apredetermined operation.

At this time, the connection force control signal CS is outputted fromthe controller 100 to the electronic control regulating valves 121A to Gassociated with the speed stage position by the speed stage positionsignal TS involved by the shift up, and the transmission 30 is set tothat speed stage position.

Subsequently, reducing the vehicle speed to shift to the hyper-slowrunning mode is desired while maintaining the power for the operation,the shift lever 111 is set to a predetermined low speed stage or a lowerspeed stage, in this embodiment, the forward first speed or second speedor the rear first speed.

Additionally, the accelerator pedal 107 is released and anon-illustrated brake is appropriately applied in order to obtain avehicle speed V which is higher than a target speed (vehicle speed) VLset in the hyper-slow running mode, for example, 1.0 km/h and lower thana relatively low predetermined speed VS, for example, 2.0 km/h.

When the accelerator pedal 107 is released, the accelerator pedal anglesignal a is outputted from the angle sensor 109 of the accelerator pedalangle detection mechanism 113, or more specifically, stepping isreleased and a signal indicative of the accelerator pedal standby modeis outputted to the controller 100.

Here, the accelerator pedal standby mode includes the state where a footis taken off from the accelerator pedal 107 so that the acceleratorpedal 107 is not stepped completely as well as the state where the footis just put on the accelerator pedal 107 but does not step on it.

Further, the speed stage position signal TS, i.e., the signalrepresenting that the shift lever 111 is set to a speed not more than apredetermined low speed stage, which is the forward first speed (F1) orsecond speed (F2) or the rear first speed (R1) in this embodiment, isoutputted from the speed stage position detection mechanism to thecontroller 100.

Moreover, the actual vehicle speed VA is calculated by the controller100 based on the output side revolution signal N2, and the controller100 makes a judgement upon whether the actual vehicle speed VA is lowerthan a predetermined speed VS.

Here, with the speed stage position being set to the forward first speed(F1) or second speed (F2) or the rear first speed (R1), when the actualvehicle speed VA is lower than the predetermined speed VS (2.0 km/h) anda signal representing that the accelerator pedal 107 is in the standbymode is inputted to the controller 100, the hyper-slow running modejudging function 104 determines the hyper-slow running mode, and thecontrol signal generating function 103 outputs the connection forcecontrol signal CS indicative of the hyper-slow running mode to theconnection force control mechanism 120.

The connection force control signal CS from the controller 100 is usedto lower the current value outputted to the electronic controlregulating valve 121 of the connection force control mechanism 120.Reduction in the current value involves a drop in the hydraulic oilpressure (working fluid pressure) to decrease the clutch connectionforce, and a slip is generated in the direction switching clutchmechanism 35A, thereby lowering the vehicle speed.

FIG. 15 shows the relationship between the hydraulic oil pressure of thedirection switching clutch mechanism 35A and the supply current value tothe proportional solenoid 121 of the electronic control regulating valve121.

In FIG. 15, the horizontal axis represents the supply current value tothe proportional solenoid 122 whilst the vertical axis represents thehydraulic oil pressure of the direction switching clutch mechanism 35A.Specific values in FIG. 15 are values given for better understanding,and they differ in accordance with each working vehicle and are thusobtained by experiments using a predetermined working vehicle.

Although the working vehicle does not start unless the hydraulic oilpressure of the direction switching clutch mechanism 35A becomes a givenmeasure or above, the hydraulic oil pressure of the direction switchingclutch mechanism 35A associated with the vehicle speed V (=0 km/h) whentrying to start is an initial pressure shown in FIG. 15, and the currentvalue 500 mA of the proportional solenoid at this time is an initialvalue.

On the other hand, in FIG. 15, with the direction switching clutchmechanism 35A being completely connected, the hydraulic oil pressure isa hold pressure, and at this time, the current value is 900 mA or above.Further, it is determined that the vehicle speed V is 1.38 km/h at theforward first speed and 1.7 k/mnh at the rear first speed.

At this moment, the vehicle speed varies depending on the forward andbackward movements because the speed decreasing ratio of thetransmission 30 differs between the forward and backward movements.

Therefore, in the hydraulic oil pressure of the direction switchingclutch mechanism 35A, the above-described initial pressure representsthe state that the direction switching clutch mechanism 35A is yet to beconnected with the force before moving the vehicle, and the holdpressure represents the complete connection. Thus, with the intermediatepressure, the direction switching clutch mechanism 35A is connectedwhile partially generating a slip and the vehicle travels at a speedlower than the vehicle speed obtained by the hold pressure.

Accordingly, it can be seen that a desired vehicle speed V suitable forthe hyper-slow running can be obtained by setting the value of thecurrent supplied to the proportional solenoid 122 so that the hydraulicoil pressure of the direction switching clutch mechanism 35A becomes theintermediate low pressure.

In FIG. 15, a hydraulic oil pressure A in the vertical axis represents ahydraulic oil pressure for obtaining a target vehicle speed VLF (=1km/h) when shifted to the forward first speed (F1) which is the forwardlow speed; a hydraulic oil pressure B, a hydraulic oil pressure forobtaining a target vehicle speed VLR (=1 km/h) when shifted to thereverse first speed (R1) which is the reverse low speed; and currentvalues 790 mA and 635 mA in this case, values of the current to besupplied to the proportional solenoid 122 during the hyper-slow running.

A flowchart of FIG. 16 shows the control procedure of the controller 100to which the above-described accelerator pedal angle signal β, the speedstage position signal TS and the output side revolution signal N2 areapplied.

In FIG. 16, when the control starts, in the controller 100, theaccelerator pedal angle signal β, the speed stage position signal TS andthe output side revolution signal N2 of the transmission 30 are read andupdated at predetermined intervals in the step 211.

Subsequently, in the step 212, a judgment is made upon whether the speedstage position is a low speed stage position, i.e., the forward firstspeed F1 or second speed F2 or the reverse fist speed R1 based on theaccelerator pedal angle signal â, the speed stage position signal TS andthe output side revolution signal N2, whether the mode is the standbymode in which the accelerator pedal 107 is not operated and whether thevehicle speed V calculated from the output side revolution signal N2 isnot more than a predetermined speed VS (speed larger than the targetvehicle speed VLF or VLR and preset to a relatively low speed), e.g.,not more than 2 km/h.

After an actual vehicle speed VA is calculated from the output siderevolution signal N2 in the step 213, a judgement is then made uponwhether the actual vehicle speed VA is within a range of a value (VL±v)obtained by adding a predetermined speed allowance v to or from thetarget vehicle speed VLF or VLR (which will be typically referred to asVL).

Here, in this embodiment, it is determined that VL is 1 km/h and v is0.1 km/h, and the above-described value (VL±v)=(1±0.1 km/h) isdetermined to be a preset deviation (dead zone).

If the actual vehicle speed VA is in a predetermined deviation, thesupply current to the proportional solenoid 122 in the pressure controlvalve 123 for controlling the direction switching clutch mechanism 35which is under control is continued without changing the currentindicated value in the step 215, and the control returns to the step 211to repeat the step 212 and the following steps.

On the other hand, if the actual vehicle speed VA is out of the presetdeviation and lower than (VL−v), the supply current to the proportionalsolenoid 122 in the pressure control valve 123 for the directionswitching clutch mechanism 35 which is under control is specified to avalue obtained by adding a correction value of a preset correction mapto the current indicated value, and the control returns to the step 211to repeat the step 212 and the following steps.

On the other hand, when the actual vehicle speed VA is out of thepredetermined deviation and is larger than (VL+v), the supply current tothe proportional solenoid 122 in the pressure control valve 123 forcontrolling the clutch mechanism 35 which is under control is todesignated to a value obtained by subtracting the correction value ofthe preset correction map from the current indicated value in the step217.

In the step 218, a judgment is then made upon whether an indicatedcurrent value I is smaller than a starting current value (initial value)IS in a flatland.

If the indicated current value I is not smaller than the initial valueIS, the supply current to the proportional solenoid 122 is againdesignated to the value obtained by subtracting the correction value ofthe preset correction map from the current indicated value in the step219 and the control then returns to the step 211 to repeat the step 212and the subsequent steps.

If the indicated current value I is smaller than the initial value IS,the current is passed through the proportional solenoid 122 of theclutch mechanism 35 on the opposite side of the driven directionswitching clutch mechanism 35, i.e., the clutch mechanism 35 on thereverse side when that on the forward side is driven or the clutchmechanism 35 on the forward side when that on the reverse side is drivenin the step 220 so as to drive in the direction opposed to that of thedriving state.

Supplying the current to the proportional solenoid 122 of the clutchmechanism 35 in the opposite direction means that the vehicle speed VAis larger than (VL+v) even though the hydraulic oil pressure of theclutch mechanism 35 on the driving side is set to the initial pressurein case of, e.g., a steep down grade. Since the vehicle speed VA can notbe further increased even if the current value on the driving side islowered to decrease the hydraulic oil pressure of the clutch mechanism35, the hydraulic oil pressure is generated in the clutch mechanism 35on the opposite side of the driving side to produce the braking force.

In the step 220, after instruction of supply of a predetermined currenttoward the clutch mechanism 35 on the opposite side, the control returnsto the step 211 to repeat the step 212 and the following steps.

FIG. 17 shows the relationship between the corrected current valuerelative to the proportional solenoid 122 of the clutch mechanism 35 onthe driving side and the vehicle speed in the above-described steps 215,216 and 217.

In FIG. 17, the horizontal axis represents the actual vehicle speed VA(km/h) and the vertical axis represents the current correction value I(mA), respectively. If the actual vehicle speed VA is within(VL±v)=(1±0.1)=(0.9 to 1.1 km/h), the current is not corrected. That isbecause, when the control is performed so as to correct only a smalldifference from the target vehicle speed VL, hunching may occur, whichcan lead to the unstable control. In other words, the predetermineddeviation set in advance functions as a so-called dead zone.

When the actual vehicle speed VA increases beyond the dead zone, i.e.,it exceeds 1.1 km/h, the correction value of the current supplied to theproportional solenoid 122 of the pressure control valve 123 in thedriven clutch mechanism 35 is linearly decreased within a predeterminedrange; it is then constantly maintained in the decreasing state; anamount of the oil supplied to the direction switching clutch mechanism35 is reduced; and the hydraulic oil pressure of the clutch mechanism 35is decreased to increase the slip, thereby lowering the vehicle speed.

On the other hand, when the actual vehicle speed VA becomes far belowthe dead zone, i.e., it becomes lower than 0.9 km/h, the currentcorrection value of the driven clutch mechanism 35 is linearly increasedin a predetermined range; it is then constantly maintained in theincreasing state; and the hydraulic oil pressure of the clutch mechanism35 is increased to decrease the slip, thereby increasing the vehiclespeed.

It is to be noted that FIG. 17 does not illustrate the operations of thesteps 218, 219 and 220 in the flowchart of FIG. 16.

Further, although the correction current is once linearlyincreased/decreased and thereafter maintained constant in FIG. 17, apredetermined correction current may be increased/decreased in thestep-like manner when the vehicle speed exceeds the dead zone asindicated by a chain double-dashed line in FIG. 17.

As indicated by the chain double-dashed line in FIG. 14, descriptionwill now be given as to the results of the embodiment in which a modechange dial and the like is provided as the speed mode setting mechanism119 to the driver seat 110.

When the speed mode setting mechanism 119 switches from the normalrunning mode to the hyper-slow running mode, a running mode signal OMinputted from the speed mode setting mechanism 119 to the controller 100is shifted to the hyper-slow running mode. As a result, a hyper-slowrunning mode judging function 104 of the controller 100 determines thatthe current mode is the hyper-slow running mode and then executes thecontrol similar to that in the step 213 and the following steps in FIG.16.

Therefore, if the speed mode setting mechanism 119 is provided, input ofthe accelerator pedal angle signal β, the speed stage position signal TSand the output side revolution signal N2 to the controller 100 isunnecessary for judging the hyper-slow running mode. However, the outputside revolution signal N2 is required for calculating the actual vehiclespeed VA to be controlled to the target speed VL.

(Advantages of the Second Embodiment)

According to this embodiment described above, the following advantagescan be obtained.

(Advantage 2-1) In this embodiment, when performing the operation in thehyper-slow running mode, the accelerator pedal angle signal â, the speedstage position signal TS and the output side revolution signal N2 of thetransmission 30 are used to check the vehicle speed V and an intent ofthe operator (the position of the shift lever 111 and the standby modeor any other mode of the accelerator pedal 107); the hyper-slow runningmode judging function 104 of the controller 100 makes a judgment uponwhether it is the hyper-slow running mode; and a predetermined directionswitching clutch mechanism 35 is controlled via the connection forcecontrol mechanism 120 in case of the hyper-slow running mode, therebycontrolling the hyper-slow running mode suitable for the situation.

In particular, the constant running state can be always obtained withrespect to the required hyper-slow running mode irrespective of the loadstate in the work or the state of the up and down grades.

In such a case, when the actual vehicle speed VA becomes larger than thetarget speed VL required in the hyper-slow running mode even if thehydraulic oil pressure of the driven direction switching clutchmechanism 35 is set to not more than the initial pressure, it ispossible to flow the current to the proportional solenoid 122 of theclutch mechanism 35 on the opposite side of the driven clutch mechanism35 to generate the braking force, which results in the wider applicationrange of the hyper-slow running mode.

Further, this embodiment can obtain (Advantage 1-1) to (Advantage 1-7)of the first embodiment. However, they are considered to be theadvantages of this embodiment, and it is determined that “hyper-slowmode” substitutes for “an operation amount of the inching pedal 114” in(Advantage 1-4) and “speed mode setting mechanism 119” substitutes for“characteristic change mechanism 118” in (Advantage 1-7).

In the second embodiment, the above-described electronic controlregulating valve 121 of FIG. 6 can be also used to similarly obtain(Advantage 1-8) and (Advantage 1-9). However, when they are applied tothis embodiment, “accelerator pedal 107” substitutes for “inching pedal114” in the above description.

(Third Embodiment)

A third embodiment according to the present invention will now bedescribed. Although this embodiment is basically similar to the secondembodiment, its control method is different.

FIG. 18 is a graph showing a different control method with respect tothe direction switching clutch mechanism 35 and illustrates an examplewhere the forward low speed clutch mechanism FL (35A) is driven.

In FIG. 18, the horizontal axis represents an actual vehicle speed VA(km/h) of a working vehicle, and an upper half portion of the verticalaxis shows a hydraulic pressure (kg/cm2) supplied to the forward lowspeed clutch mechanism FL of the direction switching clutch mechanism 35whilst a lower half of the same shows a hydraulic pressure (kg/cm2)supplied to the reverse clutch mechanism R.

A characteristic of the control according to this embodiment lies inthat the hydraulic pressure is constantly applied to the directionswitching clutch mechanism 35 on the driving side as well as that on theopposite side in the control area in the hyper-slow running mode, asshown in FIG. 18.

When the speed VA of the working vehicle is larger (faster) than apredetermined speed VS (2.0 km/h in this embodiment) greater than atarget speed VLF (1.0 km/h in this embodiment) during the usual work andthe like, the full hydraulic pressure, or more specifically, a hydraulicpressure of 23 kg/cm2 is applied to the forward low speed clutchmechanism FL. At this time, no hydraulic pressure (0 kg/cm2) is appliedto the reverse clutch mechanism R (35C).

The target speed VIF may be simply referred to as VL hereinafter.

Then, when the operator intends to set the hyper-slow running mode,he/she sets the shift lever 111 to the forward first speed, enters thestandby mode in which a foot is taken off from the accelerator pedal 107or simply put on the pedal 107 and appropriately operates the brake sothat the vehicle speed of the working vehicle VA becomes lower than apredetermined speed VS (2.0 km/h), the controller 100 judges thehyper-slow running mode.

This causes a low operating current of the first stage to be supplied tothe proportional solenoid 122 so that a low hydraulic oil pressure ofthe first stage PF1 (0.5 kg/cm2 in this embodiment) acts on the forwardlow speed clutch mechanism FL. With the operating current of the firststage, the connection force is hardly generated in the forward low speedclutch mechanism FL and the forward driving force is rarely produced inthe wheel 85.

On the other hand, an operating current of the second stage is suppliedto the proportional solenoid 122 in such a manner that a hydraulic oilpressure of the second stage PR2 (5.0 kg/cm2 in this embodiment) higherthan the hydraulic oil pressure PR1 of the first stage (0.5 kg/cm2 inthis embodiment) which will be described later acts on the reverseclutch mechanism R.

As a result, a certain degree of the connection force is generated inthe reverse clutch mechanism R, and to the wheel 85 is given the reversedriving force which acts on the working vehicle as the braking force,thereby decreasing the vehicle speed VA.

In this manner, when the vehicle speed VA is reduced to be not more thana given value which is 1.3 km/h in this embodiment, the hydraulicpressure applied to the proportional solenoid 122 is controlled so as tobe gradually decreased in accordance with reduction in the vehicle speedVA.

Then, when the vehicle speed VA becomes not more than a range obtainedby adding a fixed difference in value v=0.1 km/h to the target speed forcontrol VLF=1.0 km/h, i.e., not more than (VL+v)=1.1 km/h, the reverseclutch mechanism R does not have to act as the brake, and thefirst-stage hydraulic pressure PR1 (=0.5 kg/cm2) lower than theabove-described second-stage hydraulic pressure PR2 is applied to thereverse clutch mechanism R. Therefore, the connection force issubstantially eliminated, and the operation as the braking force is alsolost.

On the other hand, the low first-stage hydraulic pressure PR1 is appliedto the forward low speed clutch mechanism FL as described above.Accordingly, when the vehicle speed VA is within a range obtained byadding a fixed difference in value ±v=±0.1 km/h to the target speed forcontrol VL=1.0 km/h, i.e., (VL±v)=0.9 to 1.1 km/h, the low first stagehydraulic pressures PR1 and PR1 (=0.5 kg/lcm2) which are equal to eachother are applied to the both forward and reverse clutch mechanisms FLand R.

In other words, in this embodiment as mentioned above, the hydraulicpressure is constantly applied to the direction switching clutchmechanism 35 on the driving side as well as that in the opposite side inthe control area in the hyper-slow running mode.

When the vehicle speed VA becomes lower than a lower limit of the targetspeed difference for control (VL±v)=0.9 to 1.1 km/h (VA=0.9 km/h orbelow), acceleration is required, and hence the hydraulic pressureapplied to the forward low speed clutch mechanism FL is graduallyincreased. When the vehicle speed VA becomes not more than 0.7 km/h, theoperating current of the second stage is supplied to the proportionalsolenoid 122 so that the hydraulic oil pressure of the second stage PF2(7.0 kg/cm2 in this embodiment) higher than the low first stagehydraulic oil pressure PF1 (=0.5 kg/cm2) acts.

This can enables acceleration of the working vehicle.

Thereafter, in the hyper-slow running mode, a predetermined hydraulicpressure is applied to the forward low speed clutch mechanism FL or thereverse clutch mechanism R as shown in the diagram in a range where thevehicle speed in FIG. 18 is 0 km/h to 2.0 km/h, and the control iseffected in such a manner that the vehicle speed is within the controltarget range (VL±v).

Incidentally, when the reverse clutch mechanism R is driven, the forwardlow speed clutch mechanism FL side functions as the brake as opposed tothe above, and the basic control method is unchanged.

Further, the forward side second stage hydraulic pressure PLF is set tobe larger than the reverse side second stage hydraulic pressure PLR forthe same reason as described in connection with FIG. 15 that the changegear ratio of the transmission 30 is different between the forwardmovement and the reverse movement.

According to this embodiment, in the hyper-slow running mode, when thevehicle speed is lowered from a high speed, since a predeterminedpressure is applied to the direction switching clutch mechanism 35 onthe opposite side of the driving side so as to act as the braking force,the vehicle speed can be controlled to be within the target controlrange.

In particular, the sufficient adaptation to a downhill slope can be alsoobtained.

Additionally, in the hyper-slow running mode, since a hydraulic pressureis constantly applied to the forward low speed clutch mechanism FL andthe reverse clutch mechanism R in some degree, the startup of theforward low speed clutch mechanism FL and the reverse clutch mechanism Ris rapid when a new hydraulic pressure is applied to them, therebyproviding further smooth control.

That is, if the hydraulic pressure to the direction switching clutchmechanism 35 is reduced to be completely zero, a time delay is generateduntil the oil is filled in the direction switching clutch mechanism 35,but this can be eliminated in this embodiment.

It is to be noted that the control method according to this embodimentis a so-called open control method.

(Fourth Embodiment)

The control method according to a fourth embodiment of the presentinvention will now be described with reference to FIGS. 19 and 20.Although this embodiment is basically the same as the second embodiment,its control method is different.

This embodiment also exemplifies a case in which the forward low speedclutch mechanism FL (35A) is driven. That is, the forward low speedclutch mechanism FL is coupled with the speed switching clutch mechanism36 of the first speed in order to control the vehicle speed to a targetvalue of 1.0 km/h.

FIG. 19 is a view showing the relationship between the hydraulicpressure supplied to the forward low speed clutch mechanism FL and thevehicle speed VA in both the normal running mode and the hyper-slowrunning mode, in which the vertical axis represents the supply hydraulicpressure while the horizontal axis represents the vehicle speed.

FIG. 20 is a graph showing a hydraulic pressure amount obtained byincreasing or decreasing the hydraulic pressure supplied to the reverseclutch mechanism R at fixed time intervals (in accordance with eachcycle), in which the vertical axis represents the hydraulic pressure tobe increased/decreased while the horizontal axis represents the vehiclespeed.

In this embodiment, the working vehicle travels in the full or partialclutch connection (coupling) state of the forward first speed. At thistime, the forward low speed clutch mechanism FL and the speed switchingclutch mechanism 36 of the first speed are coupled by the hydraulicpressure of approximately 23 kg/cm2. On the other hand, no hydraulicpressure is applied to the reverse clutch mechanism R in this state.

In this embodiment, when the vehicle speed VA becomes lower than apredetermined speed VS (1.8 km/h in this embodiment) larger than thetarget speed VL (1.0 km/h in this embodiment) at the forward first speedor second speed or the reverse first speed and the accelerator pedal 107is in the standby mode, the control in the hyper-slow running mode isstarted.

It is to be noted that the vehicle speed VA exiting the hyper-slowrunning mode is set to 2.0 km/h and hysteresis is provided to preventhunching.

When the control in the hyper-slow running mode is started, thehydraulic pressure supplied to the forward low speed clutch mechanism FLis constantly set to 7.0 kg/cm2. This enables some slip in the forwardlow speed clutch mechanism FL.

When the hyper-slow running mode is started, comparison between theactual vehicle speed VA and the target vehicle speed VL by thecontroller 100 is simultaneously carried out. If the actual vehiclespeed VA is larger than the target vehicle speed VL (VA>VL), a fixedhydraulic pressure applied to the reverse clutch mechanism R isincreased in accordance with each one cycle. In this embodiment, onecycle is 0.5 second, and the fixed hydraulic pressure is increased by+0.25 kg/cm2 when the vehicle speed VA is larger than the target vehiclespeed VL by +0.1 km/h or more, as shown in FIG. 20.

On the other hand, if the actual vehicle speed VA is smaller than thetarget vehicle speed VL (if VA<VL), a fixed hydraulic pressure to thereverse clutch mechanism R is decreased in accordance with each onecycle. In this embodiment, as similar to the above, one cycle is 0.5second, and the fixed hydraulic pressure is decreased by −0.5 kg/cm2when the vehicle speed VA is smaller than the target vehicle speed VL by−0.1 km/h or more, as shown in FIG. 20.

Here, a range of increase in the hydraulic pressure to the reverseclutch mechanism R is set smaller than a range of decrease in the samein order not to stall by rapidly increasing the braking force.

Thereafter, the actual vehicle speed VA is compared with the targetvehicle speed VL in accordance with each one cycle (0.5 second), and thehydraulic pressure to the reverse clutch mechanism R is increased ordecreased depending on a deviation of these speeds (VA−VL) which islarger (+) or smaller (−) than 0.1 km/h, thereby controlling the vehiclespeed VA so as to be within a fixed range (VL±v) relative to the targetvehicle speed VL.

According to this embodiment, the controller 100 compares the actualvehicle speed VA with the target vehicle speed VL, and the hydraulicpressure supplied to the reverse clutch mechanism R is increased ordecreased in accordance with the deviation of these speeds (VA−VL) tocontrol the vehicle speed VA to be within the target vehicle speed range(VL±v), thus realizing the appropriate control.

In addition, since the hydraulic pressure is also supplied to thedirection switching clutch mechanism 35 on the opposite side of thedriving side, the vehicle speed can be rapidly controlled so as to bewithin the control target range, and the sufficient adaptation relativeto the down grade is possible.

Further, in this embodiment, since the hydraulic pressure is alsoconstantly applied to the forward low speed clutch mechanism FL and thereverse clutch mechanism R in the hyper-slow running mode, the startupof the direction switching clutch mechanism 35 is fast as similar to theFIG. 18 embodiment, thereby realizing the smooth control.

(Fifth Embodiment)

A fifth embodiment according to the present invention will now bedescribed with reference to FIG. 21. This embodiment is an example suchthat the hydraulic pressure supplied to the reverse clutch mechanism Rhas a degree corresponding to a deviation (VA−VL) of the target vehiclespeed VL and the actual vehicle speed VA in the hyper-slow running mode,as different from the embodiment illustrated in FIGS. 19 and 20 whereinthe hydraulic pressure supplied to the reverse clutch mechanism R has afixed value.

FIG. 21 is a graph showing the relationship between a quantity of thehydraulic pressure supplied to the reverse clutch mechanism R and thevehicle speed VA, in which the vertical axis represents the hydraulicpressure to be increased or decreased and the horizontal axis representsthe vehicle speed.

In this embodiment, when the actual vehicle speed VA deviates from thetarget vehicle speed VL beyond a predetermined range, which is 0.5 km/hor more in this embodiment and which is ½ of the value in the embodimentillustrated in FIGS. 19 and 20, the hydraulic pressure supplied to thereverse clutch mechanism R on the opposite side of the driving side isincreased or decreased in accordance with the deviation (VA−VL).

Specifically, as to increase/decrease of the hydraulic pressure appliedto the reverse clutch mechanism R, when the vehicle speed VA is higherthan the target vehicle speed VL, the hydraulic pressure is linearlyincreased by 0.2 kg/cm2 every time the vehicle speed VA is increased by0.5 km/h.

On the other hand, when the vehicle speed VA is lower than the targetvehicle speed VL, the hydraulic pressure is linearly decreased by 0.4kg/cm2 every time the vehicle speed VA is decreased by 0.5 km/h.

Here, the inclination in case of the low vehicle speed VA is sharp ascompared with that in case of the high vehicle speed VA, namely theincreasing ratio is high in order to prevent stall of the workingvehicle, as similar to the embodiment shown in FIGS. 19 and 20.

In this manner, according to the present embodiment, the advantagessimilar to the respective embodiments in FIGS. 18, 19 and 20 can beobtained, and the control can be further precisely performed.

(Modification of the Second to Fifth Embodiment)

It is to be noted that the present invention is not restricted to theforegoing respective embodiments, and modifications and improvements ina range of attaining the objects of the invention are included in thepresent invention.

For example, the connection force control mechanism 120 is notnecessarily the electronic control regulating valve 121, and a differenttype of control valve may be used. However, use of the electroniccontrol regulating valve 121 can readily obtain the connection forceaccording to the stepping angle of the accelerator pedal 107 with asimple structure.

Here, although the electronic control regulating valve 121 having thestructure of FIG. 6 does not have to be used, the small and inexpensiveapparatus can be provided if used.

Furthermore, although the clutch mechanism whose connection force iscontrolled in accordance with the stepping angle of the acceleratorpedal 107 may be the direction switching clutch mechanism 35 as well asthe speed change clutch mechanism 36 or both of them, all the speedstage positions can be advantageously controlled with a small number ofcontrol targets.

Moreover, a number of revolutions on the output side in the transmission30 does not have to be controlled so as to be within a predetermineddeviation range, hunching can be prevented from occurring if controlledin this manner.

Additionally, the speed mode setting mechanism 119 may not benecessarily provided, the hyper-slow running mode can be readily set ifprovided.

(Sixth Embodiment)

A sixth embodiment will now be described with reference to FIGS. 22 and23.

This embodiment relates to a variable power type engine which can beused in the vehicle of the foregoing respective embodiments and thepower setting method thereof.

In this embodiment, a portion surrounding the engine is similar to thatin the first embodiment, and like reference numerals denote like orcorresponding part in order to omit the tautological explanation.Different portions will be described hereunder.

FIG. 22 shows a schematic structure of this embodiment, and FIG. 23 is apower characteristic diagram showing the relationship between a numberof revolutions of the engine (rpm: horizontal axis) and an engine torque(kg/m: vertical axis) in a variable power engine of this embodiment.

In FIG. 23, description will be first given on the power characteristicrequired for the variable power engine having a fuel injector with aboost compensator.

In FIG. 23, a line A connecting diamond-shaped dots shown in thelowermost portion in the drawing by a solid line represents acharacteristic obtained when no charged pressure (boost pressure) issupplied to the boost compensator and a fuel oil consumption issubstantially fixed, and a line B connecting triangular dots shown inthe uppermost portion in the drawing by a solid line represents amaximum power characteristic brought under control by various limitswhen the fuel oil consumption is increased to give rise to the enginepower with the charged pressure being applied to the boost compensator.The limit of the maximum power is brought under control by an exhausttemperature, the durability of the engine and the like in a highrevolution region, whilst the fuel oil consumption is brought undercontrol by the exhaust gas characteristic in a low revolution area.

In case of the variable power engine, the fuel oil consumption is set soas to obtain a high power required for the vehicle between the lines Aand B, and is has a characteristic such as represented by a line Cconnecting square dots by a solid line in the drawing.

In the variable power engine whose power characteristic is set, forexample, the engine controlled by the above-described boost compensator,the characteristic of the line A or C is demonstrated in accordance withthe running state of the vehicle, the working state and others. That is,when the high power is not required in the low speed area, since thehigh power may cause the slip in the wheels and the like, the engine isdriven in the state represented by the line A in which no chargedpressure is supplied to the boost compensator. On the other hand, whenthe high power is required in the high speed area, the charged pressureis supplied to the boost compensator, and the engine is driven in thestate represented by the line C.

However, there may be a case that an intermediate power characteristicbetween the line A and the line B, e.g., a line P or a line Qrepresented by broken lines in FIG. 23 is desired.

The present invention can obtained such an intermediate powercharacteristic.

In FIG. 22, the specific structure of the sixth embodiment will bedescribed.

To the engine 10 are connected an intake pipe 11 and an exhaust pipe 12.An air cleaner 13 is connected to the uppermost stream side of theintake pipe 11, and a turbocharger 14 and an intercooler 15 are providedbetween the air cleaner 13 of the intake pipe 11 and the engine 10 fromthe upstream side to the downstream side. With this arrangement, thecleaned air passed through the air cleaner 13 is subjected to pressureapplication by the turbocharger 14 and then cooled down by theintercooler 15 to be supplied to the engine 10 in order to increase apressure charging ratio.

The fuel injector 20 and the transmission 30 are connected to the engine10.

The boost compensator 21 is attached to the fuel injector 20, and amanifold 17 branched off from the intermediate portion between theturbocharger 14 of the intake pipe 11 and the intercooler 15 isconnected to the boost compensator 21 so that the charged pressure(outlet side boost pressure) of the turbocharger 14 can be supplied. Theboost compensator 21 adjusts and controls a fuel oil consumptionrelative to the engine 10 of the fuel injector 20 in accordance with thecharged pressure of the turbocharger 14.

The transmission 30 converts output revolutions of the engine 10 into aplurality of speed stages, for example, six stages from a first speed toa sixth speed to be transmitted to wheels and the like, and the speedstage position detection mechanism (speed stage position signalgeneration mechanism) 31 is attached to the transmission 30. This speedstage position detection mechanism 31 detects which speed stage thetransmission 30 is selected to based on, e.g., the a position of theshift lever of the transmission 30 and outputs the speed stage positionsignal to the later-described pressure state switching mechanism 40.

The pressure state switching mechanism 40 includes: a fixed throttle 41which is provided to the manifold 17 and serves as a hydraulic circuitdevice having the resistive influence on the manifold 17; an auxiliarymanifold 42 which is branched off from the middle portion between thethrottle 41 in the manifold 17 and the boost compensator 21 and whosedownstream side is connected between the intake pipe 11 and the aircleaner 13 and between the intake pipe 11 and the turbocharger 14, i.e.,connected to the upstream side duct (intake pipe 11) of the turbocharger14; and switching means 50 which is provided to the auxiliary manifold42 and which can switch between a state of blocking the auxiliarymanifold 42 and another state in which the pressure of the auxiliarymanifold 42 is partially released and decreased by a pressure appliedfrom the turbocharger 14 to the manifold 17 and the obtained pressure ishigher than an atmospheric pressure.

The switching means 50 is constituted by a two-position electromagneticswitching valve 51 which blocks or communicates the middle part of theauxiliary manifold 42; and a fixed throttle 52 which is provided to theauxiliary manifold 42 on a slip stream side of the two-positionelectromagnetic switching valve 50 in the communication state of thetwo-position electromagnetic switching valve 51.

It is to be noted that the speed stage position signal is inputted fromthe above-mentioned speed stage position detection mechanism 31 to thesolenoid 53 of the two-position electromagnetic switching valve 51.

This speed stage position signal is outputted so as to block theauxiliary manifold 42 without operating the solenoid 53 of thetwo-position electromagnetic switching valve 51 when the speed stage ofthe transmission 30 is the high speed stage of, e.g., the third speed orabove, whilst it is outputted so as to operate the solenoid 53 tocommunicate the auxiliary manifold 42 when the speed stage is the lowspeed stage of, e.g., the first or second speed. When the auxiliarymanifold 42 is communicated, use of the throttle 42 to cause the chargedpressure of the turbocharger 14 to partially flow to the upstream sideof the turbocharger 14, and the pressure supplied to the boostcompensator 21 is hence decreased, thereby reducing the fuel oilconsumption of the fuel injector 20.

(Results of the Sixth Embodiment)

Results of this embodiment having the above arrangement will now bedescribed.

The air cleaned by the air cleaner 13 passes through the intake pipe 11and is charged by the turbocharger 14 and the intercooler 15 to besupplied to the engine 10.

On the other hand, although the fuel is supplied to the engine 10 by thefuel injector 20, a supply amount of this fuel is controlled by theboost compensator 21 so that the power characteristic of the engine 10is determined.

The pressure (charged pressure, boost pressure) on the outlet side ofthe turbocharger 14 via the manifold 17 is applied to the boostcompensator 21 through the throttle 41, and this pressure is changed bythe switching mode of the pressure state switching mechanism 40.

That is, when the speed stage is shifted to the low speed stage of,e.g., the second speed or a lower speed in the transmission 30 fortransmitting the power of the engine 10 to the wheels and the like, anoperation signal is sent from the speed stage position detectionmechanism 31 to the solenoid 53 of the two-position electromagneticswitching valve 51. As a result, the two-position electromagneticswitching valve 51 is switched from the state illustrated in FIG. 22 tothe upper part in the drawing to communicate the auxiliary manifold 42.Therefore, the charged pressure of the turbocharger 14 supplied to themanifold 17 partially flows out toward the upstream side of theturbocharger 14 through the throttle 52, and the pressure applied to theboost compensator 21 is decreased to a predetermined pressure.

Here, a degree of pressure reduction can be set in accordance with thethrottling state of the throttle 41 provided to the manifold 17 and thethrottle 52 provided to the auxiliary manifold 42. Further, since thethrottle 41 of the manifold 17 can throttle the charged pressure to beapplied to the boost compensator 21, the partial outflow of the chargedpressure from the auxiliary manifold 42 generates a difference inpressure between the upstream and downstream sides of the throttle 41,and a predetermined reduced pressure is applied to the boost compensator21.

When the charged pressure reduced as described above is applied to theboost compensator 21, the consumption of the fuel oil supplied from thefuel injector 20 to the engine 10 can have a small value, and the enginepower also becomes a small value. This state is the power represented bythe line P or the line Q in FIG. 23. Here, selection of the powerrepresented by the line P or the power represented by the line Q ispreviously set.

When the speed stage of the transmission 30 is set to a high speed area,e.g., the third speed or above, the ON signal outputted from the speedstage position detection mechanism 31 to the solenoid 53 of thetwo-position electromagnetic switching valve 51 in connection with thissetting is turned off, and the two-position electromagnetic switchingvalve 51 returns to the state shown in FIG. 22 by the spring force. Thiscauses the auxiliary manifold 42 to be blocked and the charged pressurefrom the auxiliary manifold 42 no longer flows out. Therefore, thecharged pressure of the turbocharger 14 acts on the boost compensator 21without being reduced and enters the high pressure state, and theconsumption of the fuel oil injected from the fuel injector 20 to theengine 10 is also increased, which leads to the high power state of theengine 10. This state corresponds to the power characteristicrepresented by the line C in FIG. 23.

Accordingly, the engine 10 has the high power in the high speed area,and hence the work is processed with the high efficiency.

(Advantages of the Sixth Embodiment)

According to the above-described embodiment, the following advantagescan be obtained.

(Advantage 6-1) In this embodiment, since the supply pressure to theboost compensator 21 can be changed to a predetermined set pressure tobe supplied by the pressure state switching mechanism 40 in accordancewith the speed stage of the transmission 30, the engine power suitablefor the content of the work can be obtained, thereby improving the workefficiency.

Specifically, the auxiliary manifold 42 is provided to the manifold 17which supplies the charged pressure of the turbocharger 14 to the boostcompensator 21, and the charged pressure is caused to partially flow out(to be discharged) or not to flow out via the auxiliary manifold 42.Therefore, a pressure which is lower than the charged pressure set bythe throttles 41 and 52 and higher than the atmospheric pressure or apressure equal to the charged pressure can be supplied to the boostcompensator 21.

Accordingly, the two-stage pressure, i.e., the charged pressure of theturbo charger 14 and a predetermined pressure which is reduced to belower than the charged pressure and higher than the atmospheric pressurecan be supplied to the boost compensator 21, and a fuel oil consumptionof the fuel injector 20, namely, the power of the engine 10 can bechanged and adjusted in two stages. Since the power can be automaticallychanged in accordance with the speed stage of the transmission 30, theappropriate running can be always enabled, thus obtaining the workingpower. In particular, since a low power can be obtained in the low speedstage and a high power can be obtained in the high speed stage, theefficient work can be carried out in the high speed stage withoutconcerning the slip of the wheel in the low speed stage.

(Advantage 6-2) Since the pressure applied to the boost compensator 21can be switched by the pressure state switching mechanism 40 constitutedby the simple hydraulic circuit device having the two-positionelectromagnetic switching valve 51 and the throttles 41 and 52, therebyinexpensively providing the apparatus.

(Advantage 6-3) Since the pressure state switching mechanism 40 can beattached to the engine 10 by adding the slight modification to themanifold 17 without adding the modification to the engine 10, the boostcompensator 21 and the like, this mechanism can be readily attached tothe construction machine, the vehicle and the like loaded with thevariable power engine system in the field.

(Advantage 6-4) Further, its maintenance is simple and no specialist isrequired. Also, replacing the throttle 52 with a counterpart having adifferent value can easily obtain the necessary engine powercharacteristic.

(Advantage 6-5) Since the hydraulic circuit device provided to themanifold 17 is the throttle 41, a pressure lower than the chargedpressure can be generated in the manifold 17 in the simple structurewhen the charged pressure partially flows out from the auxiliarymanifold 42.

(Advantage 6-6) The slip stream side of the auxiliary manifold 42 isconnected to the intake pipe 11 on the upstream side of the turbocharger14, and hence the charged pressure flowing out through the auxiliarymanifold 42 can flow back to the intake pipe 11 when the two-positionelectromagnetic switching valve 51 causes the auxiliary manifold 42 toenter the communicating state, thereby wasting no charged pressure bythe turbocharger 14.

(Modification of the Sixth Embodiment)

Various modifications of the switching means 50 in the sixth embodimentwill now be described with reference to FIGS. 24 and 25.

The embodiment shown in FIG. 24 switches the pressure supplied to theboost compensator 21 to three different pressure states.

That is, the switching means 50 shown in FIG. 24 is provided with athree-port three-position electromagnetic selector valve 54 and, amongthe three ports of the three-port three-position electromagneticselector valve 54, an A port on the upstream side is connected to theupstream side of the auxiliary manifold 42 while B and C ports on thedownstream side are connected to ducts 42A and 42B branched off from theauxiliary manifold 42 on the slip stream side, respectively. Throttles55A and 55B having different throttling conditions are provided to theseducts 42A and 42B. As to the throttling conditions of the throttles 55Aand 55B, for example, the throttle 55A has a larger amount ofthrottling, in other words, a flow rate toward the slip stream side ofthe auxiliary manifold 42 is set smaller.

The three-port three-position electromagnetic selector valve 54 hasthree positions for blocking the auxiliary manifold 42 situated in thecentral position in the drawing, for connecting the A port and the Bport in the upper position to each other and for connecting the A portand the C port in the lower position to each other, and as similar tothe first embodiment, it caused to perform the switching operation bythe solenoid 56A and the solenoid 56B activated by the speed stageposition signal from the speed stage position detection mechanism 31.

In such a structure, as shown in the drawing, when the auxiliarymanifold 42 is being connected to the central position of the three-portthree-position electromagnetic selector valve 54, since the auxiliarymanifold 42 is blocked, the charged pressure does not flow out from theauxiliary manifold 42, and the charged pressure is directly applied tothe boost compensator 21, thereby obtaining the high power of the engine10. At this time, in case of the six-stage transmission 30, thetransmission is shifted to, e.g., the fifth or sixth speed, and theswitching signal from the speed stage position detection mechanism 31 isnot applied to the both solenoids 56A and 56B.

When the speed stage of the transmission 30 is, for example, the thirdor fourth speed, the solenoid 56A is operated, and the three-portthree-position electromagnetic selector valve 54 is switched to thelower position in the drawing, thus connecting the auxiliary manifold 42on the upstream side to the auxiliary manifold 42A with the throttle 55Awhich greatly throttles. Consequently, the charged pressure partiallyflows out and a pressure decreased to be lower than the charged pressureis applied to the boost compensator 21, which results in the enginepower smaller than that described above that can be the power suitablefor the third or fourth speed.

When the speed stage of the transmission 30 is, for example the first orsecond speed, the solenoid 56B is operated, and the three-portthree-position electromagnetic selector valve 54 is switched to theupper position in the drawing, thereby connecting the auxiliary manifold42 on the upstream side to the auxiliary manifold 42B with the throttle55B which slightly throttles. This causes the charged pressure which islarger than that described above to partially flow out, and a pressurefurther decreased to be lower than the charged pressure is applied tothe boost compensator 21, thereby obtaining the further smaller enginepower which is suitable for the first and second speed.

According to the embodiment shown in FIG. 24, the following advantagescan be obtained.

(Advantage 6-7) The three-port three-position electromagnetic selectorvalve 54 can change the pressure applied to the boost compensator 21 toa total of three stages, i.e., a pressure equivalent to the chargedpressure and pressures of two stages lower than the former pressure andhigher than the atmospheric pressure, thus achieving matching of theworking speed and the power state in the improved manner.

It is to be noted that the relationship between the pressure switchingby the three-port three-position electromagnetic selector valve 54 andthe speed stage of the boost compensator 21 is not restricted to theabove-described relationship in this embodiment. That is, the presentinvention is not restricted such a setting as that the lowest power isobtained at the first and second speeds; the medium power, the third andfourth speeds; and the highest power, the fifth and sixth speeds, andthis setting can be appropriately changed in accordance with the contentof the work, such that lowest power is obtained at the first speed; themedium power, the second to fourth speeds; and the highest power, thefifth and sixth outputs. This is also true to any other embodiment.

Further, in this embodiment, since the charged pressure all flows outfrom the auxiliary manifold 42 unless the throttle 55B is provided inthe duct 42B, it is possible to set the state in such a manner thecharged pressure does not act on the boost compensator 21, namely, theatmospheric pressure acts on the same.

The embodiment shown in FIG. 25 switches the pressure supplied to theboost compensator 21 to at least four different pressures.

That is, three ducts 42C, 42D and 42E are branched off from the middleof the auxiliary manifold 42 shown in FIG. 25, and two-position selectorvalves 57A, 57B and 57C for switching between the duct blocking stateand the communication state are provided to the respective ducts 42C,42D and 42E. Further, throttles 58A and 58B having different throttlingstates are provided to the duct 42C and 42D.

According to such a structure, the following advantage can be obtained.

(Advantage 6-8) When the predetermined two-position selector valves 57Ato 57C are opened/closed in response to a signal from the speed stageposition detection mechanism 31, at least four pressure states can begenerated, and the operation status of the boost compensator 21 can bechanged in accordance with the pressure states.

Here, the phrase “at least four pressure states” is cited herein because(1) the state where all the ducts 42C, 42D and 42E are closed as shownin the drawing and (2) to (4) the states where any one of the ducts 42C,42D and 42E is communicated as well as (5) the state where the ducts 42Cand 42D are simultaneously communicated are considered, and multipletypes of the pressure states are not usually required in the actualcase.

(Seventh Embodiment)

FIG. 26 shows a seventh embodiment according to the present invention.

According to this embodiment, the pressure state switching mechanism 40provided to the manifold 17 is constituted by a selector valve and apressure reducing valve. Here, a difference between this embodiment andthe sixth embodiment is only the structure of the pressure stateswitching mechanism 40, and any other construction is the same, therebyomitting its explanation.

That is, a three-port two-position electromagnetic switching valve 61 isprovided in the middle of the manifold 17, and two ducts 17A and 17B arebranched off from the manifold 17 on the slip stream side of thetwo-position electromagnetic switching valve 61. One duct 17A directlycommunicates with the boost compensator 21 without providing anyhydraulic circuit device in the middle thereof, and the other duct 17Bis coupled with the boost compensator 21 via an internal pilot typepressure reducing valve 62 with a relief in the middle thereof.

The two-position electromagnetic switching valve 61 is provided with asolenoid 63 and a speed stage position signal is inputted from the speedstage position detection mechanism 31 to the solenoid 63. Thetwo-position electromagnetic switching valve 61 can switch the duct 17Aor the duct 17B to be connected to the manifold 17 by the speed stageposition signal. In this case, when the speed stage signal indicates,e.g., a high speed stage of the third speed or above, the manifold 17 iscaused to communicate with the duct 17A as shown in the drawing, and thecharged pressure of the turbocharger 14 is directly supplied to theboost compensator 21 (without decreasing the pressure). On the otherhand, when the speed stage signal indicates a low speed stage of thesecond speed or below, the solenoid 63 is operated to switch thetwo-position electromagnetic switching valve 61 against the spring, andthe manifold 17 is caused to communicate with the duct 17B. Further, thebehavior of the pressure reducing valve 62 is used to reduce the chargedpressure to a predetermined pressure to be supplied to the boostcompensator 21.

Therefore, in this embodiment, since the charged pressure reduced by thepressure reducing valve 62 is supplied to the boost compensator 21 inthe low speed stage of the second speed or below, the engine 10 isdriven with the low power. On the other hand, at the high speed stage ofthe third speed or above, the charged pressure can be directly supplied,thereby enabling the efficient work with the high power.

Even in the seventh embodiment having the above-described structure, theadvantages 6-1 to 6-7 similar to those in the sixth embodiment can beobtained.

That is, the pressure to be supplied to the boost compensator 31 isswitched to the multiple stages by the pressure state switchingmechanism 40, and the engine power suitable for the content of a workcan be obtained, thereby improving the work efficiency. Since this powercan be automatically changed in accordance with the speed stages of thetransmission 30, the appropriate running and operation power can beconstantly obtained. In addition, the efficient work is enabled at thehigh speed stage without concerning the slip of the wheels and the likeat the low speed stage and the mechanism can be readily attached to theconstruction machine or vehicle loaded with the variable engine systemin the field. Moreover, the maintenance is easy and any special operatoris not required. Also, when the pressure reducing valve 62 is replacedwith a pressure reducing valve having a different value, the necessaryengine power characteristic can be easily obtained. Further, such anadvantage as that the simple and inexpensive pressure state switchingmechanism 40 can be provided can be added.

(Modifications of the Seventh Embodiment)

FIGS. 27 and 28 shows different modifications of the pressure stateswitching mechanism 40 in the seventh embodiment.

The embodiment illustrated in FIG. 27 switches the pressure to besupplied to the boost compensator 21 to three different pressure states.

That is, the pressure state switching mechanism 40 of FIG. 27 isprovided with a four-port three-position electromagnetic selector valve64. Among the four ports of the four-port three-position electromagneticselector valve 64, the A port on the upstream side is connected to theupstream side of the manifold 17, and the B, C and D ports on thedownstream side are connected to the ducts 17C, 17D and 17E branched offfrom the manifold 17 on the slip stream side, respectively. Of theseducts 17C, 17D and 17E, to the ducts 17C and 17E, except the central17D, are disposed pressure reducing valves 65A and 65B set to differentpressure decreasing states. Here, in regard to the pressure reduction ofthe pressure reducing valves 65A and 65B, for example, the pressurereducing valve 65B has a larger amount of pressure reduction. In otherwords, it is set in such a manner that a pressure toward the slip streamside of the manifold 17, i.e., the supply pressure to the boostcompensator 21 becomes small.

In the drawing, the four-port three-position electromagnetic selectorvalve 64 has a central position at which the A port and the C port areconnected to each other to communicate the duct 17D; a left position atwhich the A port and the B port are connected to each other tocommunicate the duct 17C; and a right position at which the A port andthe C port are connected to each other to communicate the duct 17E. Assimilar to the second embodiment, the valve is caused to performswitching operation by the solenoids 66A and 66B activated by the speedstage position signal from the speed stage position detection mechanism31.

In such a structure, as shown in the drawing, when the manifold 17 isconnected to the central position of the four-port three-positionelectromagnetic selector valve 64, since no pressure reducing valveexists in the duct 17D, the charged pressure is directly applied to theboost compensator 21, thereby obtaining the high power of the engine 10.At this time, in case of the six-stage transmission 30, the transmission30 is shifted to, for example, the fifth or sixth speed.

When the speed stage of the transmission 30 is, for example, the thirdor fourth speed, the solenoid 66A is activated so that the four-portthree-position electromagnetic selector valve 64 is switched to theright direction in the drawing, and the manifold 17 on the upstream sideis connected to the duct 17C having the pressure reducing valve 65Ahaving a relatively small degree of pressure reduction. This ensuresreduction in the charged pressure, and a pressure decreased to be lowerthan the charged pressure is applied to the boost compensator, thusobtaining the engine power suitable for the third or fourth speedsmaller than that in the above description.

Further, when the speed stage of the transmission 30 is, for example,the first or second speed, the solenoid 66B is operated and thefour-port three-position electromagnetic selector valve 64 is switchedto the left direction in the drawing so that the manifold 17 on theupstream side is connected to the duct 17E with the pressure reducingvalve 65B having a large degree of pressure reduction. As a result, thecharged pressure is further reduced as compared with the above, and apressure obtained by further reducing the charge pressure is applied tothe boost compensator 21, thereby obtaining the smaller engine powersuitable for the first or second speed.

According to the embodiment of FIG. 27, the state of pressure applied tothe boost compensator 21 can be changed to three stages by using thefour-port three-position electromagnetic selector valve 64, and theworking speed and the power state can be matched with each other in theimproved manner.

It is to be noted that the relationship between the pressure stateswitching using the four-port three-position electromagnetic valve 64and the speed stage of the boost compensator 21 is not restricted tothat described above in this embodiment.

Since connecting the duct 17E to the upstream side of the turbocharger14 without providing the pressure reducing valve 65B in the middle partof the duct 17E can cause the charged pressure to all flow out from theduct 17E, and it is possible to set the state in which no chargedpressure acts on the boost compensator 21, i.e., the state in which theatmospheric pressure acts. At this time, an appropriate throttle or aduct resistor is provided to the duct 17E in order to prevent the entirecharged pressure from flowing out from the manifold 17 to the upstreamside of the turbocharger 14.

The embodiment shown in FIG. 28 substitutes a pair of two-positionselector valves for the four-port three-position electromagneticselector valve 64 in the embodiment illustrated in FIG. 27 so that thepressure to be supplied to the boost compensator 21 is switched to threedifferent pressure states.

That is, a pair of two-position electromagnetic switching valves 67A and67B are provided at the middle of the manifold 17 in FIG. 28, and thetwo-position electromagnetic switching valves 67A and 67B switch theducts 17F and 17G provided to the manifold 17, and the manifold 17.Further, internal pilot type pressure reducing valves each having arelief 68A and 68B which are set to different pressure reduction statesare provided to the ducts 17F and 17G, respectively.

Furthermore, solenoids 69A and 69B are provided to the two-positionelectromagnetic switching valves 67A and 67B, and the speed stageposition signal is inputted from the speed stage position detectionmechanism 31 to the solenoids 69A and 69B.

According to this structure, when the predetermined two-positionelectromagnetic switching valves 67A and 67B are operated in response tothe signal from the speed stage position detection mechanism 31, threepressure states can be generated, and the operation status of the boostcompensator 21 can be changed in accordance with the pressure states.This can be understood from the description of the foregoingembodiments.

Additionally, in the embodiment shown in FIG. 28, when any of the ducts17A and 17B is connected to the upstream side of the turbocharger 14without providing the pressure reducing valve at the middle thereof, itis possible to apply no charged pressure to the boost compensator 21 atall. At this time, an appropriate throttle or a duct resistor isprovided to the ducts 17F and 17G connected to the upstream side of theturbocharger 14 in order to prevent the entire charged pressure fromflowing out from the manifold 17 to the upstream side of theturbocharger 14, as similar to the embodiment depicted in FIG. 27.

(Another Modification of the Sixth and Seventh Embodiment)

It is to be noted that the present invention is not restricted to theforegoing embodiments and any modification or improvement within a rangefor attaining the object of the invention is included in the presentinvention.

For example, in the sixth embodiment shown in FIG. 22, the throttle 41which serves as a resistor provided to the manifold 17 may have anappropriate pipe line resistor by narrowing the duct diameter of themanifold 17.

Further, a relief valve capable of setting the discharging pressure to apredetermined pressure may be used in place of the throttle 52 providedto the auxiliary manifold 42. Although the downstream side of theauxiliary manifold 42 may be opened to the air, this structure partiallywastes the charged pressure, and the reflow to the intake pipe 11 on theupstream side of the turbocharger 14 is preferable.

Furthermore, in the present invention, the pressure state switchingmechanism 40 can be manufactured into various structures by combiningthe commercially available various hydraulic circuit devices as well asthe foregoing modifications. In brief, it is sufficient that thepressure state switching mechanism 40 can switch the pressure to besupplied to the boost compensator 21 to predetermined pressures of atleast two stages which are equal to or lower than the charged pressureof the turbocharger 14 and higher than the atmospheric pressure. Here,switching to the pressure equal to the atmospheric pressure as well asthe predetermined pressures of at least two stages can be allowed.

In addition, the expression “predetermined pressures of at least twostages equal to or lower than said charged pressure of said turbochargerand higher than a atmospheric pressure” in describing the secondembodiment the present invention intends to include one or more stagesat which the pressure is equal to or lower than the charged pressure andhigher than the atmospheric pressure and two or more stages at which thepressure is lower than the charged pressure and higher than theatmospheric pressure. Further, the phrase “one stage” in the expression“state in which a pressure of said auxiliary manifold is released and apressure to be applied from said turbocharger to said manifold isdecreased to be a predetermined pressure of at least one stage lowerthan said pressure and higher than an atmospheric pressure” in claim 21according to the present invention means that the pressure state of twoor more stages in total can be obtained if at least one stage can beacquired by pressure reduction because the boost compensator 21 has thestate of one stage at which the charged pressure is applied by the“state for closing the auxiliary manifold”.

(Eighth Embodiment)

An eighth embodiment will now be described with reference to FIGS. 29and 30.

Incidentally, FIG. 29 shows a schematic structure of this embodiment,and FIG. 30 is a graph showing the relationship between a vehicle speed(V: horizontal axis) and a rim pull (vertical axis) in a vehicle with avariable power engine such as in this embodiment.

The basic structure of this embodiment is similar to that of the sixthembodiment, and like reference numerals denote like or correspondingparts to omit the tautological explanation. Description will be given ondifferent parts.

Although not explained in the sixth embodiment, a torque converter 70with a lockup is provided between the engine 10 and the transmission 30,and the power from the engine 10 can be transmitted to the transmission30 through the torque converter 70.

In FIG. 29, an intake pipe 11 and an exhaust pipe 12 are connected tothe engine 10. An air cleaner 13 is connected to the uppermost streamside of the intake pipe 11, and a turbocharger 14 and an intercooler 15are provided to the intermediate portion between the air cleaner 13 ofthe intake pipe 11 and the engine 10 from the upstream side to thedownstream side. With such an arrangement, the air cleaned by passingthrough the air cleaner 13 is subjected to the pressure application inthe turbocharger 14 and then cooled down in the intercooler 15 to besupplied to the engine 10 so as to increase the supercharging ratio.

The fuel injector 20 and the torque converter 70 with the lockupmechanism are connected to the engine 10.

The boost compensator 21 is provided to the fuel injector 20, and themanifold 17 branched off between the turbocharger 14 of the intake pipe11 and the intercooler 15 is connected to the boost compensator 21 sothat the charged pressure (outlet side boost pressure) of theturbocharger 14 can be supplied. The boost compensator 21 adjusts andcontrols a fuel oil consumption of the fuel injector 20 relative to theengine 10 in accordance with the charged pressure of the turbocharger14.

The torque converter 70 with the lockup mechanism has a structureequivalent to that of a counterpart commercially available, and themulti-stage speed change transmission 30 is connected to the output sideof the torque converter 70. As a result, the power from the engine 10 istransmitted to the transmission 30 side while involving a slip inaccordance with a load from the transmission 30 side when the lockupmechanism of the torque converter 70 is not operated. On the other hand,when the lockup mechanism is operated, such power is directlytransmitted to the transmission 30 side without generating a slip.

Further, the torque converter 70 is provided with the lockup detectionmechanism 71 for detecting whether the lockup mechanism is operated. Thelockup detection mechanism 71 detects whether the lockup mechanism isoperated based on, for example, a position of a movable plate of theclutch constituting the lockup mechanism or a fluid pressure applied tothe clutch and outputs a lockup enabled/disabled signal to the enginepower controller 75.

The transmission 30 transmits output revolutions of the engine 10 to aplurality of speed stages, e.g., eight stages from the forward first toeighth speeds and four stages of the reverse first to fourth speeds tobe transmitted to wheels and the like, and the speed stage positiondetection mechanism (speed stage position signal generation mechanism)31 is provided to the transmission 30. This speed stage positiondetection mechanism 31 detects which speed stage the transmission isselected to based on, for example, a position of the shift lever of thetransmission 30 and outputs the speed stage position signal to theengine power controller 75.

A plurality of driving wheels 85 are coupled to the output side of thetransmission 30 via the wheel driving state switching mechanism 80. Thewheel driving state switching mechanism 80 is a mechanism for switchinga wheel 85 to be driven among a plurality of the wheels 85 which can bedriven. For example, this mechanism switches drive of only the rearwheels or the front wheels or drive of all of the front and rear wheels,i.e., all wheel drive (AWD) and this wheel driving state switchingmechanism 80 also has a structure equivalent to that of a commerciallyavailable counterpart. The wheel driving state detection mechanism 81 isprovided to the wheel driving state switching mechanism 80. The wheeldriving state detection mechanism 81 detects which is the wheel 85 to bedriven based on, for example, a position of the switching lever (orswitch) of the wheel driving state switching mechanism 80 or anengagement position of the gear of the wheel driving state switchingmechanism 80 and outputs the wheel driving state detection signal to theengine power controller 75.

The pressure state switching mechanism 40 provided to the manifold 17includes: the fixed throttle 41 as a hydraulic circuit device whichserves as a resistor to the manifold 17; the auxiliary manifold 42 whichis branched off between the throttle 41 and the boost compensator 21 inthe manifold 17 and whose downstream side is connected between the aircleaner 13 of the intake pipe 11 and the turbocharger 14, i.e.,connected to the upstream side duct (intake pipe 11) of the turbocharger14; and switching means 50 which is provided to the auxiliary manifold42 and capable of switching between the state in which the auxiliarymanifold 42 is blocked and the state in which the pressure of theauxiliary manifold 42 is partially discharged in order to obtain apredetermined pressure which is reduced to be lower than a pressureapplied from the turbocharger 14 to the manifold 17 and higher than anatmospheric pressure.

The switching means 50 is constituted by the two-positionelectromagnetic switching valve 51 for blocking or communicating themiddle of the auxiliary manifold 42 and the fixed throttle 52 providedto the auxiliary manifold 42 in the slip stream side of the two-positionelectromagnetic switching valve in the communicating state of thetwo-position electromagnetic switching valve 51.

It is to be noted that the engine power switching device 90 includes thefuel injector 20 with the boost compensator 21 and the pressure stateswitching mechanism 40 and the engine power switching device 90 switchesthe power of the engine 10 between the normal power (the normal powerside, the normal power mode) and another normal power (the normal powerside, the normal power mode, the economy mode).

The engine power controller 75 receives the lockup enabled/disabledsignal from the lockup detection mechanism 71, the speed stage positionsignal from the speed stage position detection mechanism 31 and thewheel driving state detection signal from the wheel driving statedetection mechanism 81 and outputs an engine power switching signal,which is specifically an ON/OFF signal for the solenoid 53, to thesolenoid 53 of the two-position electromagnetic switching valve 51provided to the engine power switching device 90.

Although the engine power switching signal changes the engine power inresponse to the aforesaid respective signals (the lockupenabled/disabled signal, the speed stage position signal, the wheeldriving state detection signal), it is outputted so as to block theauxiliary manifold 42 without operating the solenoid 53 of thetwo-position electromagnetic switching valve 51 when the high power isrequired. On the other hand, when the normal power which is a powerlower than the aforementioned high power is needed, the engine powerswitching signal is outputted so as to communicate the auxiliarymanifold 42 by operating the solenoid 53. When the auxiliary manifold 42is communicated, partial outflow of the charged pressure of theturbocharger 14 to the upstream side of the turbocharger 14 by thethrottle 52 lowers the pressure to be supplied to the boost compensator21, and the fuel oil consumption of the fuel injector 20 is decreased toreduce the engine power, thereby obtaining the normal power.

(Results of the Eighth Embodiment)

Results of this embodiment having the above-described structure will nowbe described.

The air cleaned by the air cleaner 13 passes through the intake pipe 11and is supercharged and cooled down by the turbocharger 14 and theintercooler 15 to be supplied to the engine 10.

Meanwhile, although the fuel is supplied to the engine 10 by the fuelinjector 20, a supply amount of this fuel is controlled by the boostcompensator 21 and the power characteristic of the engine 10 isdetermined.

The pressure (charged pressure, boost pressure) of the turbocharger 14on the output side supplied via the manifold 17 is applied to the boostcompensator 21 through the throttle 41, but this pressure variesdepending on the switching state of the pressure state switchingmechanism 40.

The switching of the pressure by the pressure state switching mechanism40 is carried out by switching the engine power switching signalsupplied from the engine power controller 75 to the solenoid 53 of thetwo-position electromagnetic switching valve 51. At this time, thetwo-position electromagnetic switching valve 51 constitutes a part ofthe switching means 50; the switching means 50, a part of the pressurestate switching mechanism 40; and the pressure state switching mechanism40, a part of the engine power switching device 90, as described above.

Therefore, it can be said that the engine power switching signal fromthe engine power controller 75 is roughly outputted to the engine powerswitching device 90, or specifically, it is outputted to the solenoid 53of the two-position electromagnetic switching valve 51.

In case of control, when the high power state (high output side) of thepower from the engine 10 is desired, the engine power controller 75transmits the high power actuation signal, which is a solenoid OFFsignal that does not actuate the solenoid 53 in this embodiment, to thesolenoid 53 of the two-position electromagnetic switching valve 51(actually, a solenoid ON signal is not transmitted). This causes thetwo-position electromagnetic switching valve 51 enters the stateillustrated in FIG. 29 by the spring force and the auxiliary manifold 42is blocked. Therefore, since the charged pressure does not flow out fromthe auxiliary manifold 42, the charged pressure of the turbocharger 14directly acts on the boost compensator 21 without being reduced, and thehigh pressure state is obtained. A consumption of the fuel oil injectedfrom the fuel injector 20 to the engine 10 is also increased so that thepower of the engine 10 becomes high.

On the other hand, when the normal power state (normal power side) ofthe power from the engine 10 is desired, the engine power controller 75transmits the normal power actuation signal, which is a solenoid ONsignal for actuating the solenoid 53 in this embodiment, to the solenoid53 of the two-position electromagnetic switching valve 51. Consequently,the two-position electromagnetic switching valve 51 is switched from thestate illustrated in FIG. 29 to the upper position in the same drawingso that the auxiliary manifold 42 is communicated. Accordingly, thecharged pressure of the turbocharger 14 supplied to the manifold 17partially flows out toward the upstream side of the turbocharger 14through the throttle 52, and the pressure applied to the boostcompensator 21 is reduced to a predetermined pressure so that the fuelsupplied from the fuel injector 20 to the engine 10 is decreased,obtaining the normal power state.

In this point, a degree of pressure reduction is set in accordance withthe throttling state of the throttle 41 provided to the manifold 17 andthe throttle 52 provided to the auxiliary manifold 42. Further, sincedisposing the throttle 41 to the manifold 17 causes the charged pressureto be throttled by the throttle 41 and applied to the boost compensator21, when the charged pressure partially flows out from the auxiliarymanifold 42, a difference in pressure is generated between the upstreamand downstream sides of the throttle 41, and a predetermined decreasedpressure is applied to the boost compensator 21.

In the control described above, switching carried out by the pressurestate switching mechanism 40 is controlled by the engine power switchingsignal from the engine power controller 75, and this control is executedin accordance with a predetermined table stored in non-illustratedstorage means such as a ROM provided in the engine power controller 75.An example of the table is shown in Table 1.

TABLE 1 High Power Mode on Neut- Forward Reverse ral 1 2 3 4 5 6 7 8 1 23 4 AWD off or no AWD Lockup x x x x ∘ ∘ ∘ ∘ ∘ x x ∘ ∘ on Lockup x x x ∘∘ ∘ ∘ ∘ ∘ x ∘ ∘ ∘ off (Torque Conv.) AWD on Lockup x x x ∘ ∘ ∘ ∘ ∘ ∘ x ∘∘ ∘ on Lockup x x ∘ ∘ ∘ ∘ ∘ ∘ ∘ x ∘ ∘ ∘ off (Torque Conv.)

That is, in Table 1, a symbol “O” represents a high power mode and asymbol “x” represents a normal mode. Table 1 is used for switchingbetween the high power mode (high power state) and the normal power mode(normal power state) based on ON/OFF of the all wheel drive (AWD) in thewheel driving state detection signal from the wheel driving statedetection mechanism 81, ON/OFF of the lockup enabled/disabled signalfrom the lockup detection mechanism 71 in ON or OFF state of the wheeldriving state detection signal, and the speed stage position signal fromthe speed stage position detection mechanism 31.

In this example, when the AWD is OFF, there is included a case of avehicle provided with no wheel driving state switching mechanism 80,i.e., a vehicle which can not switch the wheel driving state. Further,ON of the lockup enabled/disabled signal supplied from the lockupdetection mechanism 71 corresponds to a lockup enabled signal and OFF ofthe same corresponds to a lockup disabled signal which is a so-calledtorque converter mode.

Specifically, the respective state can be controlled in the followingmanner.

(1) In case of the wheel driving state detection signal fed from thewheel driving state detection mechanism 81 being set to OFF of the AWDor the vehicle provided with no wheel driving state switching mechanism80, if the lockup enabled/disabled signal from the lockup detectionmechanism 71 is ON, the high power state can be obtained when the speedstage position signal of the speed state position detection mechanism 31indicates the high speed stage position equal to or above the forwardfourth speed or the reverse third speed (F4 to F8 or R3, R4) and thenormal power state is obtained when the aforementioned signal indicatesthe lower speed stage position (F1 to F3 or R1, R2).

The control is executed in this manner because, when the high power isobtained at the low speed stage position (F1 to F3 or R1, R2) with apart of the wheels 85 being driven and the torque converter 70 with thelockup mechanism being locked up, the power is too large, and the wheels85 may generate a slip to lower the operability or wear the wheels 85.

(2) In case of the wheel driving state detection signal from the wheeldriving state detection mechanism 81 being set to OFF of the AWD or thevehicle provided with no wheel driving state switching mechanism 80, ifthe lockup enabled/disabled signal from the lockup detection mechanism71 is OFF (no lockup), i.e., the current mode is the torque convertermode in which the torque converter 70 is operated as a torque converter,the high power state can be obtained when the speed stage positionsignal of the speed stage position detection mechanism 31 indicates thehigh speed stage position equal to or above the forward third speed orthe reverse second speed (F3 to F8 or R2 to R4) and the basic normalstate can be obtained when it indicates the lower speed stage position(F1, F2 or R1).

The reason for executing such a control is similar to that of (1), andthe respective controls of (3) and (4) in the following description havethe same reason. In addition, the speed stage position for obtaining thehigh power state is lowered partly because the power loss is generatedby occurrence of some slips in the torque converter 70 in the torqueconverter mode and this power loss should be compensated and partlybecause the wheels 85 hardly generate slips in the torque converter modeeven if the high power is realized at the lower speed stage.

(3) In case of the wheel driving state detecting signal from the wheeldriving state detection mechanism 81 being set to ON of the AWD, if thelockup enabled/disabled signal of the lockup detection mechanism 71 isON, the high power state can be obtained when speed stage positionsignal of the speed stage position detection mechanism 31 indicates thehigher stage position than the forward second speed or the reversesecond speed (F3 to F8 or R2 to R4), and the normal power state can beobtained when the signal indicates the lower speed stage position (F1,F2 or R1).

(4) In case of the wheel driving state detection signal from the wheeldriving state detection mechanism 81 being set to ON of the AWD, if thelockup enabled/disabled signal from the lockup detection mechanism 71 isOFF, i.e., the current mode is the torque converter mode, the high powerstate is obtained when the speed stage position signal from the speedstage position detection mechanism 31 indicates the high speed stageposition equal to or above the forward second speed or the reversesecond sped (F2 to F8 or R2 to R4), and the normal power state isobtained when it indicates the lower speed stage position (F1 or R1).

Incidentally, in the forward first speed or the reverse first speed, thenormal power is realized irrespective of the state of the torqueconverter 70 with the lockup mechanism and the wheel driving stateswitching mechanism 80 because the high power at the start canfacilitate generation of slips in the wheels 85 regardless of theforward or reverse movement. Further, in the neutral state (N) of thetransmission basically requiring no high power, the normal power can beconstantly obtained.

Description will now be given on the running (vehicle speed) and the rimpull performance in the forward movement of the vehicle with thevariable power engine according to this embodiment illustrated in FIG.30.

FIG. 30 shows the relationship between the vehicle speed (horizontalaxis) and the rim pull (vertical axis) in a vehicle having the forwardeight-speed change gear, and a characteristic indicated by a solid linerepresents the characteristic of all the speed stages from the first toeighth speeds in the regular power (normal power) state while acharacteristic indicated by a broken line represents the characteristicof the second to eighth speeds in the high power state.

In such a vehicle, the regular work is carried out at the second speed;the light load work such as scattering a material, e.g., gravel,leveling a graveled path or snow removal, at the third to sixth speeds;and running or the lighter load work, at the seventh or eighth speed.

Here, since the high power does not lead to a problem of a slip of thewheels and the like in a light load work or running at the second speedor above, the engine power is set to the high power characteristicindicated by a broken line to perform the highly efficient work.

Referring to FIG. 30, at the second speed, in a case where the AWD is ONand the lockup enabled/disabled signal is OFF, i.e., in the torqueconverter mode in the all wheel drive, the high power mode is activated,otherwise the normal power mode is activated.

At the third speed, (1) in a case where the AWD is ON and the lockupenabled/disabled signal is ON, i.e., in the lockup mode in the all wheeldrive, or (2) in the torque converter mode irrespective of ON and OFF ofthe AWD, the high power mode is used, otherwise the normal power mode isused.

Expressing the third-speed state associated with Table 1, (1) in caseswhere the AWD is OFF and the lockup is OFF (torque converter mode), or(2) in cases where the AWD is ON irrespective of ON and OFF of thelockup, the high power mode is used.

At the fourth speed or higher speed, the high power mode is obtainedirrespective of ON and OFF of the AWD and ON and OFF of the lockup.

(Advantages of the Eighth Embodiment)

According to this embodiment having the above-described structure, thefollowing advantages can be obtained.

(Advantage 8-1) In this embodiment, since the multi-stage speed changetransmission 30 is coupled to the engine 10 via the torque converter 70with the lockup mechanism and there is provided the engine powercontroller 75 which receives the lockup enabled/disabled signal from thelockup detection mechanism 71 of the torque converter 70 with the lockupmechanism and the speed stage position signal from the speed stageposition detection mechanism 31 of the transmission 30 and outputs theengine power switching signal, the torque converter 70 enables thesmooth driving, and the lockup mechanism can improve the drivingefficiency at the high speed. Further, the optimum engine power suitablefor the content of a work can be obtained in accordance with the speedstage position of the transmission 30 and ON and OFF of the lockup ofthe torque converter 70, thus improving the working efficiency.

(Advantage 8-2) The engine power switching device 90 for switching thepower from the engine 10 in accordance with the engine power switchingsignal from the engine power controller 75 is provided with the fuelinjector 20 with the boost compensator 21 for adjusting the fuel oilconsumption in accordance with the charged pressure of the turbocharger14 and the pressure state switching mechanism 40 for switching thesupply pressure to the boost compensator 21, and hence the chargedpressure of the turbocharger 14 can be utilized to readily switch thepower from the variable power engine 10. Here, switching of the chargedpressure can be facilitated by the pressure state switching mechanism40.

(Advantage 8-3) The wheel driving state detection signal is alsoinputted from the wheel driving state detection mechanism 81 of thewheel driving state switching mechanism 80 to the engine powercontroller 75, and the optimum engine power control can be thus effectedin accordance with the driving state of the wheels 85.

(Advantage 8-4) Specifically, since the auxiliary manifold 42 isprovided to the manifold 17 for supplying the charged pressure of theturbocharger 14 to the boost compensator 21 and the charged pressure ispartially flowed out (discharged) or not flowed out through theauxiliary manifold 42, a predetermined pressure which is lower then thecharged pressure and higher than the atmospheric pressure and which isset by the throttles 41 and 52 or a pressure equivalent to the chargedpressure can be supplied to the boost compensator 21.

(Advantage 8-5) Therefore, the pressures of two stages, i.e., thecharged pressure of the turbo charger 14 and the predetermined pressurewhich is reduced to be lower than the charged pressure and higher thanthe atmospheric pressure can be supplied to the boost compensator 21,and the fuel oil consumption of the fuel injector 20 which can be thepower of the engine 10 can be changed and adjusted in two stages. Sincethe power can be automatically changed by the engine power switchingsignal from the engine power controller 75 in accordance with the speedstage of the transmission 30, ON and OFF of the lockup of the torqueconverter 70 with the lockup mechanism and the wheel driving state ofthe wheel driving state switching mechanism 80, and hence theappropriate running can be always enabled, thus obtaining the workingpower. In particular, since the low power is obtained at the low speedstage and the high power can be obtained at the high speed stage, theefficient work can be carried out at the high speed stage without aconcern over a slip of the wheels and the like at the low speed stage.

(Advantage 8-6) Since the pressure to be supplied to the boostcompensator 21 can be switched by the pressure state switching mechanism40 constituted by the simple hydraulic circuit device which is thetwo-position electromagnetic switching valve 51 and the throttles 41 and52, thereby inexpensively providing the apparatus.

(Advantage 8-7) Since the pressure state switching mechanism 40 can beattached to the engine 10 by slightly adapting the manifold 17 withoutaltering the engine 10 or the boost compensator 21 and the like, it canbe readily attached to the construction machine or vehicle loaded withthe variable power engine system.

(Advantage 8-8) Further, its maintenance is simple and a specialoperator is not required. Also, replacing the throttle 52 with acounterpart having a different value can easily obtain the necessaryengine power characteristic.

(Advantage 8-9) Since the hydraulic circuit device provided to themanifold 17 is the throttle 41, a pressure lower than the chargedpressure can be generated in the manifold 17 with a simple structurewhen the charged pressure partially flows out from the auxiliarymanifold 42.

(Advantage 8-10) Since the slip stream side of the auxiliary manifold 42is connected to the intake pipe 11 on the upstream side of theturbocharger 14, the charged pressure flowing out through the auxiliarymanifold 42 can flow back to the intake pipe 11 when communicating theauxiliary manifold 42 by the two-position electromagnetic switchingvalve 51, and the charged pressure produced by the turbocharger 14 isnot wasted.

(Modification of the Eight Embodiment)

It is to be noted that the present invention is not restricted to theforegoing embodiments and any modification and improvement can beincluded in the present invention as far as the object of the inventioncan be attained.

For example, the engine power switching device 90 is not restricted tothe structure of the above-described embodiments, and it may be anapparatus realized by an electronic governor described in the backgroundart, an apparatus disclosed in U.S. Pat. No. 4785778, an apparatus whichcan easily turn on/off the charged pressure of the turbocharger suppliedto the boost compensator by the electromagnetic valve or any otherapparatus. In brief, an apparatus capable of switching the engine powerto multiple stages can suffice.

Further, in the present invention, the wheel driving state switchingmechanism 80 does not have to be provided, and the present invention canbe also applied to a vehicle in which only a predetermined wheel 85 isconstantly driven. In such a case, the control by the engine powercontroller 75 is the same as the control when the AWD is OFF. In thiscase, the present invention can be applied to a vehicle using a crawlerin place of the wheels 85.

When providing the wheel driving state switching mechanism 80, it is notrestricted to a mechanism using all the wheels 85 as driving wheels, anda mechanism capable of increasing/decreasing a number of driving wheelsmay be used. In this case, the control by the engine power controller 75can be considered to be equal to the control when the AWD is ON.

Moreover, the speed stage position for switching the high power stateand the normal power state is not restricted to those in Table 1, and itcan be appropriately changed in accordance with the engine powercharacteristic, usage of the vehicle or any other situation. However,when the high power mode is used at the higher speed stage than apredetermined speed stage and the lockup is ON in this state, the speedstage to be switched is set to a higher speed stage. Further, even if anumber of driving wheels is increased, the speed stage to be switched isset to a higher speed stage.

Additionally, a number of speed stage positions of the multi-stage speedchange transmission 30 is not restricted to the forward eight stages andthe reverse four stages in the foregoing embodiments, it may beappropriately determined.

Further, the pressure state switching mechanism 40 is not limited to theabove-mentioned embodiments, and any other configuration may be adopted.For example, a plurality of the auxiliary manifolds 42 of the switchingmeans 50 may be provided, and the two-position selector valve and thethrottle may be provided to each of these ducts. When a value of thesethrottles is changed, different pressures of three or more stages ratherthan two stages can be supplied to the boost compensator 21, and theengine power can be changed to the multiple-stage power suitable forthese pressures.

Furthermore, the selector valve of the switching means 50 may be athree-position selector valve, and a plurality of ducts are provided onthe slip stream side of the three-position selector valve. Throttleshaving different values may be provided to a plurality of the ducts toobtain pressures of the multiple stages as similar to the above. Also,various kinds of known hydraulic circuit devices may be used toconstitute the pressure state switching mechanism 40. For example, thetwo-position selector valve and the pressure reducing valve may be usedto configure the pressure state switching mechanism 40, and a so-calledgeneral fluid pressure circuit technique can be applied.

What is claimed is:
 1. A variable power engine comprising: aturbocharger for supplying a charged pressure to an engine; a fuelinjector with a boost compensator for adjusting a fuel oil consumptionin accordance with said charged pressure of said turbocharger; atransmission for changing output revolutions of said engine to aplurality of speed stages; a manifold for supplying said chargedpressure of said turbocharger to said boost compensator; a pressurestate switching mechanism which is provided to said manifold and capableof switching a pressure to be supplied to said boost compensator topredetermined pressures of at least two stages equal to or lower thansaid charged pressure of said turbocharger and higher than anatmospheric pressure; and a speed stage position detection mechanism fordetecting a speed stage position of said transmission to supply a speedstage position signal to said pressure state switching mechanism, saidpressure state switching mechanism being caused to perform pressurestate switching operation by said speed stage position signal suppliedfrom said speed stage position detection mechanism.
 2. The variablepower engine according to claim 1, wherein said pressure state switchingmechanism comprises: a hydraulic circuit device which is provided tosaid manifold and demonstrates a resistance action or a pressuredecreasing action with respect to said manifold; an auxiliary manifoldbranched off between said hydraulic circuit device and said boostcompensator in said manifold; and switching means which is provided tosaid auxiliary manifold and capable of switching between a state inwhich said auxiliary manifold is blocked and a state in which a pressureof said auxiliary manifold is released and a pressure to be applied fromsaid turbocharger to said manifold is decreased to be a predeterminedpressure of at least one stage lower than said pressure and higher thanan atmospheric pressure.
 3. The variable power engine according to claim2, wherein said hydraulic circuit device provided to said manifold isconstituted by a throttle.
 4. The variable power engine according toclaim 2, wherein said switching means provided to said auxiliarymanifold is constituted by a two-position selector valve capable ofswitching between a duct blocking state and a communicating state and athrottle provided to a duct on a slip stream side of said two-positionselector valve in said communicating state of said two-position selectorvalve.
 5. The variable power engine according to claim 2, wherein saidslip stream side of said auxiliary manifold is connected to a duct on anupstream side of said turbocharger.
 6. The variable power engineaccording to claim 1, wherein said pressure state switching mechanism iscaused to switch to said state for blocking said auxiliary manifold whensaid speed stage position signal supplied from said speed stage positiondetection mechanism is a high speed stage position signal.
 7. A variablepower engine comprising: a turbocharger for supplying a charged pressureto an engine; a fuel injector with a boost compensator for adjusting afuel oil consumption in accordance with said charged pressure of saidturbocharger; a transmission for changing output revolutions of saidengine to a plurality of speed stages; a manifold for supplying saidcharged pressure of said turbocharger to said boost compensator; apressure state switching mechanism which is provided to said manifoldand capable of switching a pressure to be supplied to said boostcompensator to a pressure equal to said charged pressure of saidturbocharger and a predetermined pressure of at least one stage lowerthan said charged pressure and higher than an atmospheric pressure; anda speed stage position detection mechanism for detecting a speed stageposition of said transmission to supply a speed stage position signal tosaid pressure state switching mechanism, said pressure state switchingmechanism being caused to perform pressure state switching operation bysaid speed stage position signal supplied from said speed stage positiondetection mechanism.
 8. The variable power engine according to claim 7,wherein said pressures state switching mechanism provided to saidmanifold is constituted by a two-position selector valve capable ofswitching said communicating state of said manifold to two positions anda pressure reducing valve provided to at least one duct on a slip streamside of said two-position selector valve.
 9. A power setting method fora variable power engine, said variable power engine comprising: aturbocharger for supplying a charged pressure to an engine; a fuelinjector with a boost compensator for adjusting a fuel oil consumptionin accordance with said charged pressure of said turbocharger; and atransmission for changing output revolutions of said engine to aplurality of speed stages, wherein said charged pressure of said turbocharger is supplied to said boost compensator as predetermined pressuresof at least two stages equal to or lower than said charged pressure ofsaid turbocharger and higher than an atmospheric pressure to set anengine power in accordance with a speed change position of saidtransmission.
 10. The power setting method of a variable power engineaccording to claim 9, wherein a pressure equal to said charged pressureof said turbocharger is supplied to said boost compensator to set anengine power when a speed stage position of said transmission is a highspeed stage position.